QforttEU Inioeraita Siibtatg
3tl{aca. Hew ^atk
B'OUGHT WITH THE INCOME OF THE
SAGE ENDOWMENT FUND
THE GIFT OF
HENRY W. SAGE
1891
Cornell University Library
TJ 795.P98
Diesel engine design,
3 1924 004 406 116
The original of tliis book is in
tlie Cornell University Library.
There are no known copyright restrictions in
the United States on the use of the text.
http://www.archive.org/details/cu31924004406116
DIESEL ENGINE DESIGN
DIESEL ENGINE
DESIGN
BY
H. F. P. PURDAY, B.Sc, A.C.G.I.
NEW YORK
D. VAN NOSTRAND COMPANY
Printed in Great Britain
PREFACE
This book is based, on about twelve years' experience of
Diesel Engines, mainly from the drawing-office point of view,
and is intended to present an account of the main considera-
tions which control the design of these engines.
The author ventures to hope that, in addition to designers
and draughtsmen, to whom such a book as this is most naturally
addressed, there may be other classes of readers— for example,
Diesel Engine users and technical students — ^to whom the
following pages may be of interest.
The text deals mainly with general principles as exemplified
by examples of good modern practice, and it has not been
possible to notice every constructional novelty. Apology is
perhaps called for on account of the omission of any special
treatment of the stepped piston and the opposed piston types
of engine. These, however, are the specialities of a compara-
tively limited number of manufacturers, and have been very
fully described and illustrated in the technical press.
The existence in its fourth edition of Chalkley's well-known
book on The Diesel Engine for Land and Marine Purposes has
enabled the present writer to proceed to details with a minimum
of preliminary discussion. A number of references to other
books and papers have been inserted in order to avoid, so far
as possible, overlapping with other sources of information.
The author has pleasure in acknowledging his indebtedness
to : Mr. P. H. Smith (who has at all times placed his unique
experience of Diesel Engines at the disposal of the author) for
several corrections and suggestions ; to Mr. L. Johnson, M.A.
(Cantab.), for his very careful and patient revision of the
proofs ; to the author's wife for assistance with the manuscript
and for compiling the index.
H. F. P. P.
CONTENTS
CHAPTER I
FIBST PBINCIPLBS
The Diesel principle — Compression pressure and temperature —
The four stroke cycle — The two stroke cycle — Types of Diesel
Engines ......... Page 1
CHAPTER II
THEBMAL EFPICIENCY
Calorific value of fuel — Laws of gases— An ideal Diesel Engine- —
Fuel consumption of ideal and real Diesel Engines — Mechanical
efficiency — ^Mechanical losses — Efficient combustion — Entropy
diagrams — Specific heat of gaseous mixtures . . .16
CHAPTER III
EXHAtrST, SUCTION AND SCAVENGE
Renewal of the charge — Flow of gases through orifices — Suction
stroke of four stroke engine — Scavenging of two stroke engine
— Exhaust of two stroke engine ...... 37
CHAPTER IV
THE PBINCIPLB OF SIMILITUDE
Properties of similar engines — Weights of Diesel Engines —
Determination of bore and stroke — Piston speed- — Mean
indicated pressure . . . . . . . .53
CHAPTER V
OBANK-SHAFTS
Materials — General construction- — Order of firing — Details of con-
struction — Strength calculations — Twisting moment diagrams 65
viii DIESEL ENGINE DESIGN
CHAPTER VI
rLY-WHBELS
The functions of a fly-wheel — Fly-wheel effeet — Twisting moment
diagrams — Degree of uniformity — Momentary governing —
Alternators in parallel — Torsional oscillations and critical
speeds — Moment of inertia — Types of fly-wheels — Strength
calculations ........ Page 97
CHAPTER VII
FRAMBWOEK
"A" frame type of framework — Crank-case t3^e — Trestle type —
Staybolt type — Main bearings — Lubrication — Strength of bed-
plate — Cam-shaft drive — Bedplate sections— Strength of "A"
frames — Design of crank-cases — Machining the framework . 119
CHAPTER VIII
CYLINDBBS AND COVBBS
Types of cylinders — Liners — Jackets — Cylinder lubrication —
Types of cylinder cover — Strength of cylinder covers . . 141
CHAPTER IX
ETJNNING GEAR
Trunk pistons — Shape of piston crown — Proportions of trunk
pistons — Gudgeon-pins — Water-cooled trunk pistons — Pistons
for crosshead engines — Systems of water cooling — Piston rods
— Crossheads and guides — Guide pressure diagrams — Connect-
ing rods — Big ends — Small ends — Strength of connecting rods
Big end bolts — Indicating gears ..... 161
CHAPTER X
FUBL OIL SYSTBM
External fuel system — Fuel system on engine — Fuel pumps —
Governors — Governor diagram — Fuel injection valves — Open
type — Augsburg type — Swedish type — Burmeister type —
Flame plates and pulverisers — ^Design of fuel valve casing and
details .... ^ .... . 190
CHAPTER XI
AIR AND EXHAUST SYSTEM
Air suction pipes — Suction valves — Exhaust valves — Exhaust
lifting devices — Proportions of air and exhaust valves and
casings — Exhaust valve springs — Inertia of valves and valve
gear — Spring formulae — Exhaust piping — Silencers — Scav-
engers — Scavenge valves — Controlled scavenge ports — Ex-
haust systems for two stroke engines ..... 220
CONTENTS ix
CHAPTER XII
COMPEESSED AIB SYSTEM
Functions of the air-blast — Capacities of blast-air compressors —
Number of stages — Compressor drives — Construction of com-
pressor cylinders — ^Volumetric efficiency — Stage ratios — Trans-
mission of heat through plates — Air reservoirs — Air-bottle
valves — Blast pipe system — Starting pipe system — Starting
valves — Air motors ....... Page 243
CHAPTER XIII
VALVE GEAB,
Cams — Cam rollers — Valve levers — Fulcrum shafts for valve
levers — Push rods — Cam-shafts — Cam-shaft bearings and
brackets — Cam-shaft drives — Reversing gears — Sliding cam-
shaft type — Twin cam-shaft type — Twin roller type — Selective
wedge type — Two stroke reversing gears^Moving roller type
— ^Manoeuvring gears — Interlocking gears — Hand controls . 263
Index 297
LIST OF ILLUSTRATIONS
No. PAGE
1. Curve connecting compression pressure and temperature for
various values of " n " 4
2. Indicator card from four stroke Diesel Engine .... 6
3. Light spring card from four stroke Diesel Engine .... 7
4. Valve-setting diagram for four stroke engine 8
5. Valve-setting diagram for two stroke engine 9
6. Diagram showing simple port scavenge 10
7. Diagram showing controlled port scavenge 12
8. Ideal indicator diagram 20
9. Fuel consumption curves for ideal and actual engines ... 24
10. Entropy temperature diagram 30
11. Calibrated indicator diagram 32
12. Diagram illustrating flow of gases through orifices ... 38
13. Curve relating suction-pressure and throat velocity ... 40
14. Curve relating scavenge pressure and throat velocity . . 41
15. Valve area diagram 42
16. Diagram showing position of piston at various crank angles . . 43
17. Valve and port area diagram for two stroke engine ... 44
18. Light spring indicator card from two stroke engine ... 52
19. Three cylinder slow speed land engine 56
20. Weight per B.H.P. of slow speed engines 57
21. Built up crank-shaft 65
22. Solid shaft for two-cylinder engine .66
23. Orders of firing for fotir stroke engines 68
24. Orders of firing for two stroke engines 68
25. Section through ring lubricator 70
26. Crank fitted with centrifugal ring lubricator 70
27. Longitudinal and transverse oil holes in crank-shaft . . . 71
28. Diagonal oil holes in crank webs . 71
29,30,31. Various shapes of crank webs 71
32, 33, 34. Methods of securing balance weights 72
35. Cast iron flanged coupling ........ 73
36. Solid forged coupling 73
37. 38. Compressor cranks .74
39. Diagram showing loads on four-crank shaft 80
40. Diagram showing deflections of imiform cantilever . . . 82
41. Construction for deflected shape of beam ... .82
42. Diagram for obtaining coefficients c™, e^, g„ etc. . 83
43. Loads on four-crank shaft — ^No. 1 firing 84
44. Loads on four-crank shaft — ^No. 2 firing 86
45. Diagram showing positions of connecting rod 90
46. Calibrated indicator card 91
47. Cylinder pressure curve on crank-angle base 93
48. Twisting moment ctirves for four cylinder engine .... 93
49. Portion of twisting moment curve 98
xi
xii DIESEL ENGINE DESIGN
No. PAOB
50. Resultant twisting moment curve for one cylinder, four stroke
engine 101
51. Resultant twisting moment curve for two cylinder, four stroke
engine 102
52. Resultant twisting moment curve for four cylinder, four stroke
engine 102
53. Resultant twisting moment curve for eight cylinder, four stroke
engine 102
54. Resultant twisting moment curve for three cylinder, four stroke
engine 102
55. Resultant twisting moment curve for six cylinder, four stroke
engine 102
56. Resultant twisting moment cxirve for one cylinder, two stroke
engine 103
57. Resultant twisting moment curve for two cylinder, two stroke
engine 103
58. Resultant twisting moment curve for four cylinder, two stroke
engine 103
59. Resultant twisting moment curve for eight cylinder, two stroke
engine 103
60. Resultant twisting moment curve for three cylinder, two stroke
engine 103
61. Resultant twisting moment curve for six cylinder, two stroke
engine 103
62. Part of resultant twisting moment curve for three cylinder, four
stroke engine, 20" bore X 32" stroke 106
63. 64. Diagrams for obtaining angular variation 106
65. Uniform shaft and fly-wheel 109
66. Diagram showing disposition of masses in a shaft system . . Ill
67. Construction for finding radius of gyration of a soUd of revolution 113
68. Solid disc fly-wheel 114
69. Disc fly-wheel with loose centre 114
70. Large fly-wheel in two pieces 114
71. Fly-wheel 115
72. Fly-wheel 117
73. " A " type of framework for four stroke engine . . . .119
74. " A " type of framework for two stroke engine .... 120
75. Four stroke " A " column with tie-rod 121
76. Four stroke " A " column with tie-rod and crosshead guide . 121
77. " A " column for foiu: stroke marine engine 120
78. " A " column for two stroke engine, with crosshead . . . 120
79. Crank-case framework for four stroke engine 122
80. Steel staybolts for crank-case 122
81. Crank-case framework for two stroke engine 122
82. Crank-case framework for crosshead engine (low tjrpe) . . . 123
83. Crank-case framework for crosshead engine (high tsrpe) . . 123
84. Trestle type of framework .% 124
85. Staybolt framework for trunk engine "... . 125
86. Staybolt framework for crosshead engine 125
87. Oil-catcher for main bearings 128
88. Forced lubricated bearings 128
89. Main bearing girder . . 129
90. Lower spiral drive . . . .131
91. Bedplate sections . . 132
92. " A " column 135
93. Crank-case ....... .... 137
94. Types of crank-case construction 138
LIST OF ILLUSTRATIONS xiii
No. FACE
95. Cylinder liner jacket and cover 141
96. Integral liner and jacket 141
97. Integral liner jacket and cover . . ■ 142
98. Liner for four stroke trunk engine 143
99. Liner for four stroke crosshead engine 143
100. Upper flange of cylinder jacket 146
101. Cylinder jacket for four stroke engine 147
102. Cylinder jacket for two stroke engine 149
103. Cylinder jacket for two stroke engine 149
104. Cylinder jacket for two stroke engine . . ... 149
105. Cylinder lubricating fittings 150-2
106. Old type of cylinder cover 153
107. Modem type of cylinder cover 154
108. Cylinder cover with separate top plate 155
109. Water tube fitting 156
110. Water inlet in side of cover 156
111. Section of cylinder cover 157
112. Cylinder cover for two stroke engine 158
113. Cylinder cover for two stroke engine 159
114. Cylinder cover for two stroke engine 160
115. Types of piston crown 162
116. Diagram showing temperature gradient in piston . . . .163
, 117. Water-cooled pistons 163
118. Loose piston top ...... . . . 164
119. Device to obviate cracking of the piston crown . . . .164
120. Diagram showing proportions of trunk pistons . . .164
121. Means of locating piston rings - 165
122. Tapering of trunk pistons 166
123. Reduction of piston surface round gudgeon-pin bosses . . .166
124. Alternative forms of gudgeon-pins 167
125. SUt system of gudgeon-pin lubrication 168
126. Piston crown 168
127. Piston cooling systems 169
128. Pistons for four stroke crosshead engines ..._.. 171
129. Piston cooling systems 171
130. Piston for two stroke crosshead engine 172
131. Piston for two stroke crosshead engine 172
132. Piston for two stroke crosshead engine 172
133. Lower end of piston rod 174
134. Lower end of piston rod 174
135. Crosshead guide block 175
136. Diagram showing relation between guide pressure and twisting
moment 175
137. Forms of connecting rod big end 177
138. Cast iron big end " brasses " 178
139. Spigoted big end 178
140. Ring for relieving big end bolts of shear stress .... 178
141. Types of connecting rod small ends 179
142. Top end of marine connecting rod 180
143. Small end with drag links 181
144. Curves showing variation of transverse stress in connecting rods,
due to inertia ... 181
145. Diagram showing deviation of connecting rod thrust, due to
journal friction 182
146. Buckled connecting rod 182
147. Resultant thrust when friction at journal =0 183
148. Resultant thrust when friction at crank-pin =0 . . . .183
xiv DIESEL ENGINE DESIGN
No. ^AGE
149. Outline of conneoting rod 185
150. Diagram showing unequal pull in big end bolts .... 188
151. Indicator gears 189
152. External fuel system for land engine 190
153. Diagrammatic arrangement of fuel pumps and governor . . 193
154. Diagrammatic arrangement of fuel pump, governor, and fuel
distributors 193
165. Fuel distributor 194
156. High pressure pipe union 194
157. Fuel pimip for markie engine 195
158. Fuel pump for one cylinder 196
159. Fuel pump ~ ... 197
160. Fuelpump 198
161. Horizontal fuel pump 199
162. Horizontal fuel pump 201
163. Diagram showing hand control of fuel pump valves . . . 201
164. Control handle for fuel pump 201
165. Plunger connections 202
166. Plunger packing 202
167. Fuel pump valves 202
168. Hand plunger 203
169. Diagram of fuel pump delivery 204
170. Governor 205
171. Diagram of governor controlling forces 207
172. Speed varying device 208
173. Speed varying device 208
174. Arrangement of governor and fuel pump gear 209
175. Arrangement of governor and fuel pump gear 209
176. Arrangement of governor and fuel piunp gear 210
177. Open type of fuel valve 211
178. Augsburg type of fuel valve 212
179. End of fuel valve 213
180. Sleeve type of pulveriser 214
181. Swedish type of fuel valve 215
182. Burmeister fuel valve 216
183-86. Lever arrangements for fuel valve operation .... 217
187. Air suction strainer 220
188. Suction pipe common to a number of cylinders .... 221
189. Suction valve casing 222
190. Cast iron exhaust valve heads 224
191. Water-cooled exhaust valve casing 224
192. Valve spindle guides 225
193. Valve casing with renewable seat 226
194. Exhaust lifting devices (hand) 227
195. Exhaust lifting devices (mechanical) 228
196. Exhaust valve proportions 229
197. Valve directly operated by caA 230
198. Velocity and acceleration curves for tangent cam .... 232
199. Diagram of valve and levers 233
200. Exhaust system for land engine 235
201. Exhaust system for land engine 236
202. Cast iron silencer 237
203. Scavenge pump driven off crank-shaft 238
204. Lever drive for scavenge pump 238
205. Combined scavenge and L.P. compressor cylinders . . . 238
206. Scavenge valve 240
207. Valves for controlling scavenge porta 241
LIST OF ILLUSTRATIONS xv
No. PAQE
208. Silencer 242
209. H.P. piston-rings for air compressor 246
210. Heat flow diagram 250
211. Diagrammatic arrangement of H.P. air-bottles .... 252
212. Valve for bottle-head 253
213. Air-bottle with head screwed to neck 254
214. Blast air pipe for land engine 255
215. Blast shut-off valve 255
216. Starting pipe for fom: cylinder land engine 256
217. Flange for starting pipe 256
218. Tee-piece for starting pipe 256
219. Starting valve 257
220. Starting valve 259
221. Starting valve 259
222. Diagrammatic arrangement of Burmeister starting valve . 259
223. Pneumatic cylinder with oil bufier cylinder 260
224. Rotary air motor 261
225. Exhaust cams 263
226. Double cam for reversing engine 265
227. Fuel cam-piece 265
228. Tangent exhaust cam profile 265
229. Smooth cam profile 265
230. Construction for cam profile 265
231. Profile based on clearance circle 265
232. Starting cam 266
233. Valve roller 266
234. Cast iron valve roller 266
235. Fulcrum shaft and valve levers 267
236. Diagram of starting, neutral and running positions . . . 268
237. Sections of valve levers 269
238. Forked ends of levers 269
239. Tappet screws 270
240. Swivel tappet 270
241. Dimensions of levers and fulcrum spindle 271
242. Outline of exhaust lever 272
243. Section of lever 272
244. Plan of lever 272
245. Push-rod ends 273
246. Stepped cam-shaft and cam-trough 274
247. Cam-shaft with separate bearings 274
248. Cam-trough 275
249. Spiral drive for cam-shaft 276
250. Vertical shaft couplings 277
251. Spiral and bevel drive for cam-shaft 278
262. Spin- drive for cam-shaft 278
253. Spiu' and coupling rod drive for cam-shaft 278
254. Suspended gear-box 279
255. Diagram for spiral gears 280
256. Reversing gear for four stroke engine 281
257. Reversing gear for four stroke engine 282
258. Mechanism for sliding cam-shaft 282
259. Double cam-shaft reversing gear 283
260. Twin roller reversing gear 284
261. Reversing gear for starting valves 285
262. Selecting wedge type of reversing gear 285
263. Valve settings for two stroke angina 286
264. Reversing gear for two stroke engine 287
xvi DIESEL ENGINE DESIGN
No. PAGE
265. Reversing gear for two stroke engine 287
266. Reversing gear for two stroke engine 289
267. Reversing gear for two stroke engine 290
268. Reversing gear for two stroke engine 291
269. Diagrammatic arrangement of manoeuvring gear .... 293
270. Interlocking gear for parallel shafts . . .... 295
271. Interlocking gear for shafts at right angles 295
DIESEL ENGINE DESIGN
CHAPTER I
FIRST PRINCIPLES
The Diesel Principle. — ^The characteristic feature of the
Diesel Engine is the injection of oil fuel into air which has been
previously compressed by the rising of a piston to a pressure
corresponding to a temperature sufficiently high to ensure
immediate ignition of the fuel.
In the course of the pioneer experiments by which the
commercial practicability of this engine was demonstrated, it
was found advantageous to effect the injection of the fuel by a
blast of air, and this feature was retained in all Diesel Engines
until the lapse of the original patents.
At the present date there exists a class of high-compression
oil engines operating on the Diesel principle, in which the
injection of oil is effected by mechanical means without the
assistance of an air blast. The design of these engines presents
a variety of problems differing materially from those which
arise in the design of Diesel Engines as defined below. Further-
more, the use for war purposes of one of the most conspicuous
members of this class of engine prohibits anything like a satis-
factory discussion of these so-called " solid injection " engines.
The well-known " surface ignition," " hot bulb," or " hot
plate " engines form a very numerous class by themselves and
have in the past been misnamed " semi-Diesel " engines. The
cycles on which they operate and the principles underlying
their design differ so widely from those relating to Diesel
Engines proper that they also fall outside the scope of this
work.
The features which characterise the true Diesel Engine, in
the correct use of the term, are now understood to be the
following : —
DIESEL ENGINE DESIGN
(1) Compression sufficient to produce the temperature
requisite for spontaneous combustion of the fuel.
(2) Injection of fuel by a blast of compressed air.
(3) A maximum cycle pressure (attained during combustion)
not greatly exceeding the compression pressure, i.e.
absence of pronounced explosive effect.
Item 3 is deliberately worded somewhat broadly as the
shape of a Diesel indicator card is subject to considerable
variation under difEerent conditions of load, blast air pressure,
fuel valve adjustment, etc.
In the earlier days of Diesel Engine construction the square
top indicator card, showing a period of combustion at constant
pressure, was considered the ideal to aim at. It has since been
found that a card having a more peaked top is usually associ-
ated with better fuel consumptions. When tar oil is used as
fuel the square top card appears to be almost out of the
question.
It should further be remembered that the existence of a
period of combustion at constant pressure is no guarantee that
all the combustion takes place at that pressure. This ideal is
never realised. Combustion probably proceeds slowly well
after half stroke, even under the most favourable conditions.
Compression Pressure. — The height to which compression
is carried is governed by the following considerations : —
(1) The attainment of the requisite temperature.
(2) The attainment of a desirable degree of efficiency.
(3) Mechanical considerations.
Considerations of temperature for ignition fix the lower
limit of compression at somewhere in the neighbourhood of
400 lb. per sq. in. The temperature actually attained depends
on the initial temperature of the intaken air and the heat lost
to the jacket during compression, so it is clear that the temper-
ature attained on the first few strokes of the engine will be
considerably lower than the'^alue it assumes after the .engine
has been firing consecutively for some time.
As regards efficiency, it is well known that increasing the
degree of compression beyond certain limits does not very
materially increase even the theoretical efficiency.
In practice the compression most usually adopted is about
500-550 lb. per sq. in. for four stroke engines. For two stroke
engines the compression is frequently in the neighbourhood of
FIRST PRINCIPLES
600 lb. per sq. in. or over, owing to the fact that the charge of
air delivered by the scavenge pump may itself be at a pressure
slightly above atmospheric. At first sight it might be thought
that this initial compression might be considered as a first
stage in the temperature rise of the charge of air ; but
apparently compression in the scavenge pump is not so
effective in raising the temperature as compression in the main
cylinder. The mechanical considerations which limit the
compression are numerous, and some are mentioned below.
Higher compression involves : —
(1) Heavier load per sq. in. of the piston and necessitates
massive construction of all the main parts.
(2) More highly compressed air for injection and consequently
increased trouble with the air compressor, and its
valves particularly.
(3) Increased wear of cylinder liners due to increased
pressure behind the piston rings.
Compression Temperature. — ^With a compression of 500 lb.
per sq. in. in a fair-sized four cycle cylinder working under
full load conditions the compression temperature is about
1200° F. On starting the engine from a cold state the com-
pression pressure and temperature are considerably lower
owing to the cold state of the cylinder walls and the piston
crown.
In addition to this the injection of cold blast air with the
fuel in the proportion of about 1 lb. of blast air to 12 lb. of
suction air still further reduces the temperature apart from
the probability that the blast air has momentarily a local
cooling effect in the zone of combustion.
The middle curve (Fig. 1) shows graphically the connection
between the compression temperature and compression pressure
on the assumptions that : —
(1) The initial temperature of the intaken air is 212° F.
(2) That the exponent in the equation PV''=const. is 1-35.
These assumptions correspond approximately to the condi-
tions obtaining with a fully loaded engine of fair size — say an
18" cylinder.
The noteworthy point about this curve is the slowing down
of the rate of increase of temperature with pressure as the
latter increases. Expressed mathematically dT diminishes as
P increases. dP
DIESEL ENGINE DESIGN
The Four Stroke Cycle. — The well-known four stroke cycle
consists briefly of : —
(1) The Suction Stroke.
(2) The Compression Stroke.
(3) The Combustion and Expansion Stroke.
(4) The Exhaust Stroke.
These are considered in detail below.
■noo
100 200 300 too 500 600 100
Compression-Lbs per Sq. In Gauj«Pressure ,
Fig. 1.
Suction Stroke. — If the engine crank is considered to be
at its inner dead centre and just about to begin the suction
stroke, the suction valve is already slightly open. In steam
engine parlance it has a slight lead. At the same time the
exhaust valve, which has been previously closing on the
exhaust stroke, has not yet come on its seat. The result of
this state of afEairs is that the rapidly moving exhaust gases
create a partial vacuum in the combustion space and induce a
flow of air through the suction valve, thus tending to scavenge
out exhaust gases which would otherwise remain in the
cylinder.
As the piston descends its velocity increases and reaches a
maximum in the neighbourhood of half stroke. At the same
time the suction valve is being lifted further off its seat and
FIRST PRINCIPLES
attains its maximum opening also in the neighbourhood of
half stroke. The lower half of the suction stroke is accom-
panied by a more or less gradual closing of the suction valve,
which, however, is not allowed to come on its seat until the
crank has passed the lower dead centre by about 20°. At the
moment when the crank is passing the lower dead centre the
induced air is passing through the restricted opening of the
rapidly closing suction valve with considerable velocity and an
appreciable duration of time must elapse before the upward
movement of the piston can effect a reversal of the direction of
flow through the suction valve. It will be clear from the above
that owing to the effect of inertia more air will be taken into
the cylinder in the manner described than by allowing the
suction valve to come on its seat exactly at the bottom dead
centre. The exact point at which the suction valve should
close is doubtless capable of approximate calculation, but is
usually fixed in accordance with current practice or test-bed
experiments.
Compression Stroke. — The piston now rises on its up
stroke and compresses the air to about 500 lb. per sq. in. The
clearance volume necessary for this compression being about
8% of the stroke volume. During the compression the ternper-
ature rises and a certain amount of heat is lost to the cylinder
walls and cylinder cover. The final compression temperature
is in the neighbourhood of 1200° F.
Combustion and Expansion Stroke.— At the upper dead
centre, or slightly previous thereto, the injection valve opens
and fuel oil is driven into the cylinder and starts burning im-
mediately. The actual point at which the fuel enters the
cylinder is not qtiite certain, as there is inevitably some lag
between the opening of the injection valve and the entrance of
fuel. The point at which the fuel valve starts to open, as
determined by a method described below, varies from about
3° (slow speed engines) to 14° (high speed engines). The
method of determining the point of opening of the fuel valve
is as follows : —
With the engine at rest, air at about 100 lb. pressure is
turned on to the injection valve and then communication with
the blast air-bottle is cut off to prevent unnecessary waste of
air and the possibility of the engine turning under the impulse
of the air which is subsequently admitted to the cylinder. The
indicator cook is now opened and the engine slowly barred
DIESEL ENGINE DESIGN
round by hand until the air is heard to enter the cylinder by
placing the ear to the indicator cock. The position of the
engine when this occurs is the nearest possible approximation
to the true point of opening, assuming the operation has been
oarefuUy done.
The duration of the fuel valve opening is usually about 48°,
and in the majority of engines is fixed for aU loads. It is
evident that at light load the opening is longer than necessary,
and in some designs the duration of opening is regulated by the
governor in accordance with the load.
The combustion is by no means complete when the fuel valve
closes, and usually continues in some measure well past the
half stroke of the engine. This is known as " after burning,"
Fio. 2.
and takes place with the very best engines in the best state of
adjustment. Exaggerated after burning is the surest sign of
misadjustment, and makes itself apparent by abnormally high
terminal pressure at the point at which the exhaust valve
opens, and is readily detected on an indicator card by com-
parison with that taken from an engine in good adjustment.
As will be shown later, the presence of " after burning " is
most clearly seen on an Entropy Diagram.
Expansion continues acc(mipanied by loss of heat to the
cylinder walls until the exhaust valve opens.
Exhaust Stroke. — The exhaust valve opens about 50°
before the bottom dead centre, in order that the exhaust gases
may efEect a rapid escape and reduce the back pressure on the
exhaust stroke. The pressure in the cylinder when the exhaust
valve starts to open is about 40 lb. per sq. in. with an engine
working with a mean indicated pressure of 100 lb. per sq. in.
The temperature of the exhaust gases at this point is some-
FIRST PRINCIPLES
where in the neighbourhood of about 1600° F. and the velocity
is consequently very high.
The pressure falls nearly to atmospheric shortly after the
bottom dead centre has been passed and the back pressure
during the remainder of the exhaust stroke should not be more
than about 1 lb. per sq. in., or less. Excessive back pressure
may arise from : —
(1) Insufficient diameter or lift or late opening of exhaust
valve.
(2) Exhaust pipe too small in diameter.
(3) Obstructions or sharp bends in the exhaust pipe or
silencer.
(4) Interference by another cylinder exhausting into the
same pipe.
It is interesting to note that owing to the higher velocity of
air at high temperature per unit pressure difference the back
pressure is more at light load than at full load.
Indicator Cards. — Figs. 2 and 3 show typical indicator
cards taken with a heavy and a light spring respectively.
The latter is particularly useful for investigating the pro-
cesses of suction and exhaust.
It is to be observed that in Fig. 3 the compression is seen to
start at a point which is indistinguishable from the bottom
dead centre, thus indicating a volumetric efficiency of practi-
cally 100%. This is to be regarded as a normal state of affairs,
obtainable with both high speed and low speed engines. The
volumetric efficiencies of internal combustion engines are
frequently quoted at figures varying between about 95% for
slow speed engines to 80% for high speed engines. The former
DIESEL ENGINE DESIGN
figure is reasonable, but the latter can only be due either to
imperfect design (or adjustment) of the engine or to erroneous
indicator cards. The use of too weak a spring in the indicator
may lead to a diagram showing not more than 60% volumetric
eflftciency, owing to the inertia of the indicator piston, etc.
Consequently, fairly stifE springs are to be preferred.
Valve Setting Diagram. — Fig. 4 is a typical valve setting
diagram for a four stroke engine, and shows the points relative
to the dead centres at which the various valves open and close.
Fuel Valve Open§_
Suction Valve Opens, ^^
'xhaust Valve Closes
fuel Valve Closes
Suctfan Valve
Closes
Exhaust Valve
Opens
Fig. 4.
The Two Stroke Cycle. — As its name implies, the two
stroke cycle is completed in one revolution of the engine. The
revolution may roughly be divided into three nearly equal
parts : —
(1) Combustion and Expansion.
(2) Exhaust and Scavenged
(3) Compression.
The exhaust and scavenge take place when the piston is
near the bottom dead centre, and consequently only very small
portions of the expansion and compression strokes are lost, in
spite of the fact that nearly 120° of the crank revolution are
occupied with exhaust and scavenge. This point is clearly
seen on reference to Fig. 5.
FIRST PRINCIPLES
9
Exhaust Period. — The exhaust starts when the piston un-
covers slots in the cylinder wall. The point at which this
happens is different in different designs of engine, an average
being about 15% of the stroke before bottom dead centre.
The exhaust ports are usually of large area, and consequently
the pressure falls to atmospheric very rapidly. The period
required for this process naturally depends on the port area
and the piston speed, and average figures are about 20° to 30°.
It is well to dwell carefully on the state of affairs at this
point.
Fuel Yslve Closes
Admission of
Scairenge Air Ceases
Exhaust Ports Close
Exhaust Port s Open
Scavenge Air is Admitted
Valve Setting
2 Stroke
Diag ram for
Engine .
Fic. 6.
During exhaust the cylinder pressure has fallen from about
55 to about 15 lb. per sq. in. absolute, and there is no reason
to suppose that the remaining exhaust gases have fallen greatly
in temperature. (Given adiabatic expansion, the fall in
absolute temperature is less than 20%.) The conclusions are,
therefore : —
(1) Something like 50% by weight of the gases have effected
their escape.
(2) The remaining gases are rarefied compared with atmo-
spheric air.
Scavenge Period. — The scavenge air is admitted by ports
or valves (or both), and the instant at which admission starts
is timed to coincide with that at which the cylinder contents
10
DIESEL ENGINE DESIGN
attain appreciably the same presstire as the scavenge air, or a
trifle less. The incoming scavenge air is supposed to sweep the
remaining exhaust gas before it and so fill the cylinder vfdth a
charge of pure air by the time the piston has covered the
exhaust slots on the up stroke. Actually certain processes
take place which do not enter into the ideal programme.
Some of these are : —
(1) A certain amount of mixing between the incoming
scavenge air and the retreating exhaust gases.
( 2) Short circuiting of scavenge air to the exhaust pipe before
all the exhaust gas has been expelled.
The effects of both these processes are minimised by provid-
ing a large excess of scavenge air. The figure adopted for the
ratio of scavenger volume to cylinder volume was about 1 -4 in
the earlier designs, but later experience points to the advis-
ability of increasing this figure to about 1-8.
There are a number of different systems in use for admit-
ting scavenge air, and some of these are discussed below.
Simple Port Scavenge. — In this system the scavenge air
is admitted by means of ports in the cylinder liner opposite a
Exhaust
Ports
Scavenge
Ports
Diagrammatic Arrangement of
Simpleton Scavenge
Fit
Fio. 6.
row of similar ports for the exhaust (see Fig. 6), the piston
top being provided with a projection to deflect the scavenge air
to the top of the cylinder. This system is simple but possesses
some disadvantages which are enumerated below.
(1) The scavenge air slots have to be made shorter than the
exhaust slots in order that the cylinder pressure may fall to
FIRST PRINCIPLES 11
the same value as the scavenge air pressure before the piston
begins to uncover the scavenge slots. This entails the latter
being covered by the piston on its upward stroke before the
exhaust ports are covered, and consequently the pressure at
the beginning of compression can barely exceed the pressure
in the exhaust pipe. There is also a possibility of exhaust gases
working back into the cylinder.
(2) The projection on the top of the piston necessitates an
irregularly shaped cylinder cover in order to provide a suitable
shape for the combustion space.
Engines provided with this system of scavenge are only
suited for a relatively low mean indicated pressure of about
80 lb. per sq. in.
Cylinder Cover Valve Scavenge. — In this system the
scavenge air is admitted by means of one to four valves located
in the cylinder cover, and avoids some of the disadvantage of
the simple port scavenge.
By allowing the scavenge valves to close after the exhaust
ports have been covered by the piston, the cylinder may
become filled with air at scavenge pressure before compression
starts, and consequently such a cylinder is capable of develop-
ing a higher mean effective pressure. The greatest drawback
to this system is the complication of the cylinder cover, and
this appears to be rather serious. Strenuous efforts are being
made to design suitable cylinder covers for this type of two
stroke engine, but with the exception of one or two designs,
which are now in the experimental stage, most of them appear
to have a comparatively short life.
Valve Controlled Port Scavenge. — This system (usually
associated with the name of Messrs. Sulzer Bros.) appears to
combine most of the advantages of both with the disadvantages
of neither of the above systems.
The air ports, or a certain number of the air ports, are bo
situated that they are uncovered before the exhaust ports, but
are controlled by a valve of the double beat, piston, or other
type, in such a way that communication does not exist between
the cylinder and the scavenge pipe until the exhaust ports
have been uncovered for a sufficient period to allow the cylinder
pressure to fall to or below the scavenge pressme. On the up-
ward stroke the controlling valve remains open, so that the
cylinder is in communication with the scavenge pipe until the
scavenge slots are covered by the piston. (See Fig. 7.)
12
DIESEL ENGINE DESIGN
Engines controlled on this principle are at present the most
successful of the large two stroke Diesel Engines.
Amongst small or medium powered two stroke engines the
simple port scavenge principle is the favourite.
Sca\/eng^
Ports
Diagrammatic Arrangement of
2 Stroke Cylinder with Controlled
Port Scavenge,
Fig. 7.
Types of Diesel Engines. — The existing types of Diesel
Engines can be divided into groups in various ways, according
as they are : —
(1) Stationary or Marine.
(2) Four Cycle or Two Cycle.
(3) Slow Speed or High Speed.
(4) Vertical or Horizontal.
(5) Single Acting or Double Acting.
It suffices here to describe shortly the outstanding features of
the commonest types in commercial use.
Four Stroke Slow Speed Stationary Engines. — ^There are
probably more Diesel Engines falling under this heading
than any other. The usual Mrangement is a vertical engine
having one to six cylinders provided with trunk pistons. The
main features are (1) a massive cast-iron bedplate made in one
or two pieces, the bottom being formed in the shape of a tray
to catch any lubricating oil which is thrown out of the bearings
or drips from the cylinders. (2) Massive columns forming the
cylinder jacket into which the liners are pressed. (3) A heavy
fly-wheel of large diameter, necessitated by the low speed of
revolution.
FIRST PRINCIPLES 13
These are perhaps the most rehable Diesel Engines hitherto
constructed, and differ but little from the machines made by
the Maschinenfabrik Augsburg in the very earliest days of
Diesel Engine manufacture. For economy in fuel oil, and
lubricating oil consumption these engines are unsurpassed, and
the massiveness of their construction secures for them a long,
useful life.
Somewhat similar engines are made in horizontal form, in
accordance with popular gas engine practice.
The horizontal engine is obviously suitable where head room
is limited, and it is a trifle cheaper to manufacture than the
vertical design. The latter, however, is still the general
favourite, and is made in very large numbers in standard sizes,
varying from 8 to 200 B.H.P. per cylinder.
Four Stroke High Speed Stationary Engines. — ^The high
speed Diesel Engine was introduced in order to cheapen manu-
facture and to provide a more compact prime mover for use in
limited spaces. Forced lubrication is usually adopted, and an
enclosed crank-case is therefore provided. The fly-wheel is of
course much smaller and lighter than that of a slow speed
engine of the same power. In other respects the design follows
fairly closely that of the slow speed engine only that additional
care has to be exercised in the design of those details in which
inertia plays an important part. The fuel consumption is
practically the same as that of a slow speed engine, but the
lubricating oil consumption is usually higher and the
useful life of the engine somewhat shorter.
Two Stroke Slow Speed Stationary Engines. — ^These are
built in sizes varying from about 200 to 700 B.H.P. per cylinder.
Their chief advantage over the four cycle equivalent is reduced
cost and size. The fuel consumption is slightly higher than
that of the four stroke engines of the same power and the cost
of upkeep would also appear to be slightly greater. These
engines are usually fitted with a cross-head, and this enables
forced lubrication to be used without an unduly high con-
sumption of lubricating oil. The use of cross-heads is also
becoming increasingly common with four cycle engines, and
this appears to be a move in the direction of increased
reliability.
Other Types of Stationary Engines. — Horizontal two
stroke Diesel Engines have been made in small numbers.
Examples of double-acting horizontal engines are also to be
14 DIESEL ENGINE DESIGN
found, but the three types described above appear to fulfil most
requirements for land service.
Four Stroke Marine Engines. — Most of the successful
marine Diesel Engines are of this type. Cross-head and guides
are usually fitted and forced lubrication is preferred. The
piston speed is usually moderate, being about 700 to 850 per
minute. In Messrs. Burmeister, Wain's well-known design the
bedplate and columns form an enclosed box construction of
great rigidity and strength. The total enclosure of the crank-
pase is also conducive to economy of lubricating oil. In other
respects the engines follow land practice very closely in almost
every essential point of design. It has been frequently con-
tended in the past that the failure of certain marine Diesel
Engines has been almost entirely due to lack of knowledge of
the requirements of marine service. The actual facts of the
case appear to be that almost every complete failure of a
marine Diesel Engine has been due to causes which would have
led to the same result on land. Under the category of failures
must be placed a number of large two stroke Diesel Engines,
which have been installed on board ship before they had proved
their reliability on land. The actual modifications required to
convert a successful land design into a successful marine design
are comparatively trivial. When once the essential difficulties
of Diesel Engine construction have been successfully sur-
mounted the adaptation to the requirements of marine service
is a small matter in comparison.
Two Stroke Marine Engines. — ^As indicated in the pre-
vious paragraph, there have been many failures with large two
stroke engines. In cases where the manufacturers have had
no experience of marine work, they appear to have been un-
duly influenced in favour of accepted marine traditions.
Where the engine has been constructed by marine steam
engineers, under licence from a Diesel Engine manufacturing
firm, the result has been similax. The machines resulting from
this fusion of ideas have generally resembled marine steam
engines with Diesel Engine cylinders. The columns, guides
and bedplate are scarcely to be distinguished from those of a
marine steam engine. The open crank-pit, adopted out of
deference to steam engine practice, appears to lead to an
excessive consumption of lubricating oil, and renders the
economical use of forced lubrication impossible, besides
facilitating the ingress of dirt. Another feature borrowed from
FIRST PRINCIPLES 15
steam practice consists of the rocking levers driven off the
cross-head for working the scavenge air pumps. These have
not always been quite successful. These and other circum-
stances all seem to point to the advisability of developing the
marine Diesel on lines a little more independent of steam engine
precedent. At present the two stroke marine Diesel Engine,
of large power, must still be considered to be in the experi-
mental stage (without prejudice to two or three isolated cases
of large two stroke Diesel Engines at present operating success-
fully). For moderate powers, up to about 500 H.P., two stroke
engines have been fairly widely and successfully used.
High Speed Marine Engines. — ^A large number of designs
of both four stroke and two stroke high speed marine Diesel
Engines have been described in the technical press of recent
years.
Many of these differ but little (except in the provision of a
reversing gear and other details) from high speed land engines.
Others, particularly of the two stroke type, have been
developed along strongly individual lines for some special
purpose, such as the propulsion of submarines.
To this latter class belong the stepped piston engines, in
which an extension of the main piston serves at once as a
cross-head guide and a single acting scavenge pump piston.
With one or two exceptions, these engines appear difficult to
dismantle, and are not likely to become popular until this
objection is removed.
Literature.— Chalkley, A. P., "The Diesel Engine."— Con-
stable.
WeUs and Taylor, " Diesel or Slow Combustion Oil Engines."
— Crosby Lockwood.
Supine, Bremner and Richardson, "Diesel Engines." — Griffin.
Milton, J. T., " Diesel Engines for Sea-going Vessels." —
Inst. N. A., April 6th, 1911.
Milton, J. T., " The Marine Diesel Engine."— Inst. N. A.,
April 2nd, 1914.
Knudson, I., " Performance on Service of Motor Ship
8uecia."—ln6t. N. A., June 27th, 1913.
CHAPTER II
THERMAL EFFICIENCY
The overall thermal efficiency of a heat engine is the ratio
of the useful work performed to the mechanical equivalent
of the heat supplied during a given period of working.
Problem : What is the thermal efficiency of a Diesel Engine
which consumes 0-4 lb. of fuel per brake horse-power hour ?
To solve this problem two things require to be known : —
(1) How much heat 1 lb. of fuel gives out on combustion,
i.e. the calorific value of the fuel.
(2) How much mechanical work is equivalent to a given
quantity of heat.
The calorific value of different qualities of liquid fuel varies
from about 15,300 (Mexican crude) to about 19,300 (Galician
crude) British Thermal Units per lb. For calculations and
comparison of test results fuel consumptions are usually
reduced to their equivalents at a calorific value of 18,000
B.T.U. per lb. One B.T.U. (the amount of heat required to
raise the temperature of 1 lb. of water 1° F.) is equivalent to
778 ft. lb. This is Joules' equivalent. Therefore, since 1 H.P.
hour = 1,980,000 ft. lb., the required thermal efficiency is equal
to:—
1,980,000 ^
0-4x18,000x778 '
From this it is seen that roughly one -third of the heat
supplied has been converted into useful work. It is within the
province of thermodynamics to determine what proportion of
the heat loss is theoretically unavoidable, and to what extent
the performance of an actual engine approximates to that of
an ideal engine working on the same cycle of operations. It is
not proposed to give here more than a brief summary of the
physical laws relating to the behaviour of air under the in-
16
THERMAL EFFICIENCY 17
fluence of pressure and temperature, which form the basis of
thermodynamic investigations.
Pressure, Volume and Temperature of Air. — The relation
between these three quantities is expressed by the formula : —
P.V.=53-2 xw xT — (1) where P= Pressure in lb. per sq. ft.
abs.
V=Volume in cubic ft.
w=Weight in lb. of the
quantity of air under con-
sideration.
T=: Temperature in degrees
abs. P.
This relation holds good for any condition of temperature
and pressure, and for a specified weight of air, given the values
of any two of the quantities represented by capital letters,
the third can be calculated.
Example : Find the volume of 1 lb. of air at atmospheric
pressure and 60° F. In this case P=14-7 lb. per sq. in. abs.,
T = 60 + 461=521° abs. F., and w = l.
Hence: V=53-2x621 ,^ , . .^
-^-T-f, — =-- = 131 cub. ft.
147 xl44
Isothermal Expansion and Compression. — If the tempera-
ture remains constant during compression or expansion the
process is said to be isothermal, and the value of T in equation
(I) becomes a constant quantity. Thence for isothermal pro-
cesses equation (1) becomes: P.V.=constant (2).
If 1 lb. of air is under consideration the value of the constant
is equal to 63-2 times the absolute temperature.
Work done during Isothermal Compression. — If P^ and
Vi represent the pressure and volume before compression,
and Pj and Vg the same quantities after expansion, then : —
V
Work done=const. log^ ^' (3),
the constant being that of equation (2). It should be borne
in mind (though the bare fact can only be stated here) that
the internal energy of a gas depends on its temperature only,
regardless of the pressure. It therefore follows that aU the
work done in isothermal compression must pass away as heat
through the walls of the containing vessel, and for this reason
18 DIESEL ENGINE DESIGN
^_ — -
isothermal processes are not attainable in practice, though
they may be approximated to by slow compression in cylinders
arranged for rapid conduction of heat.
Specific Heat at Constant Volume. — ^If 1 lb. of air is
heated in a confined space, i.e. at constant volume, 0-169
B.T.U. are required to raise the temperatxu-e by 1° F. This
then is the specific heat of air at constant volume. For many
purposes it is near enough to consider the specific heat as
constant, though actually its value increases sUghtly as the
temperature increases. The amount by which the internal
energy of 1 lb. of air increases as the temperature rises is there-
fore : —
0-169 (Ta— Tj) where (T2—Ti)=the increase of tempera-
ture.
Specific Heat at Constant Pressure. — In this case, on the
other hand, work is done by expansion if heat is being added,
and by compression if heat is being discharged ; consequently
the specific heat at constant pressure exceeds that at constant
volume by the equivalent of the work done. The specific
heat at constant pressure is 0-238 B.T.U. per lb. per degree
Fahrenheit.
Adiabatic Expansion and Compression. — Expansion or
compression unaccompanied by the transfer of heat to or from
the air is termed Adiabatic. It should be noted that the air
may lose heat by doing external work or gain heat by having
work done on it by the application of external force during an
adiabatic process ; but this heat comes into existence or passes
out of existence within the air itself and does not pass through
the walls of the containing vessel.
The following relation holds good between the pressure and
the volume during an adiabatic process : —
P V° =constant (4) where n =ratio of the two specific
heats, \<fe. — 1-41.
Owing to the conductivity of the cylinder walls adiabatic
compression and expansion are not to be obtained in practice,
but it is frequently possible to express the actual relation
between pressure and volume by means of equation (4) if suit-
able values are chosen for " n." A little consideration will
shew that for compression with some loss of heat the value of
" n " will be less than 1-41, and for expansion with some loss
THERMAL EFFICIENCY 19
of heat greater than 1'41. An expansion or compression in
which the relation between P and V is expressed by equation
(4) is known as a Poljrtropic Process.
Work done on Poly tropic Expansion of Air, — The work
done is the integral of the pressure with respect to the volume,
and if expressed in ft. lb. is given by : —
-^^P]Ti^z£iX_2 = 53-2 ('^i~'^a) (5)
n— 1 n— 1
for 1 lb. of air, where the suffixes 1 and 2 have their usual
significance in denoting the state before and after expansion ;
the work done during a corresponding compression between
the same pressures is of course the same.
Temperature Change during Poly tropic Expansion.— By
combining equations (1) and (2) the following is obtained : —
, J:=(r:)"Kvir-'»>
With the information supplied by the above six equations it
is possible to construct the Indicator Diagram corresponding
to an Ideal Diesel Engine, in which all defects of combustion
and heat losses to the cylinder walls are supposed to be
eliminated.
An Ideal Diesel Engine. — Before proceeding further, it is
necessary to define what is to be understood by the term
ideal engine for the purposes of this investigation. In the
first place, frictional losses and all leakage are eliminated and
compression and expansion are supposed to take place adiabat-
ically (n= 1 -41). The specific heats are supposed to be constant
and the same whether for air or a mixture of air or exhaust
gases. On the other hand, the provision of compressed air for
the purpose of fuel injection will be recognised, and conse-
quently the machine will have a mechanical efficiency less than
unity.
The compression in the air compressor will be regarded as
isothermal and the compressor itself free of all mechanical or
other losses.
It will also be supposed that all the exhaust gas, including
that contained in the clearance space, is ejected on the exhaust
stroke and that every suction stroke fills the cyHnder with air
at 14-7 lb. per sq. in. abs. and 521° abs. F. The stroke volume
is taken for convenience of calculation at 100 cub. ft.
20
DIESEL ENGINE DESIGN
The Ideal Indicator Card. — ^The indicator card correspond-
ing to this ideal engine is now readily constructed, and is shown
in Fig. 8.
Point A denotes a volume of air equal to 100 cub. ft. plus the
clearance space, which has to be calculated.
Line AB represents adiabatic compression from 14 -7 lb.
per sq. in. to 514-7 per sq. in. abs.
n.'^
Volume in Cubic Feet. '
/ Fig. 8.
Line BC represents increases of volume at constant pressure
due to addition of heat by combustion of fuel.
Line CD represents adiabatic expansion of the products of
combustion and DA the fall in pressure due to exhaust release.
It is not necessary to deal with the exhaust and suction strokes,
as these are supposed to take place at atmospheric pressure.
Determination of Clearance Value. —
41
From equation (4).
P« 514-7
m-
100
This determines Points B and A.
Construction of Line AB. — ^This is effected by calculating
the pressure at various points of the stroke in accordance with
the following schedule : —
Pa ~
14-7 ""
\35/f4i
12-45
= 11-45 .-.
Vb = 8-75 cub. ft
THERMAL EFFICIENCY
21
1
2
3
4
s
e
7
Per
cent of
stroke.
V-
cubic feet.
Va
V
1..;='
(«X1.«.
AntUoK
(5)
P=(6)X14-7
lb. per
sa. in.
108-75
1-00
0-000
0-000
1-00
14-7
20
88-75
1-23
0-099
0-127
1-34
19-7
40
68-75
1-58
0-199
0-280
1-90
27-9
60
48-75
2-23
0-348
0-490
3-09
45-5
80
28-75
3-78
0-579
0-815
6-53
95-5
90
18-75
5-79
0-763
1-075
11-89
174-5
95
13-75
7-90
0-898
1-265
18-41
270-6
100
8-75
12-45
1-095
1-545
35-08
514-7
This simple calculation is given in detail, as it is typical.
Determination of Point c. — ^The value of V^ depends of
course on the amount of heat added. The efEect of the oil itself
in increasing the weight of the working fluid will be ignored.
The blast air, however, will be taken into consideration and
assumed to be equivalent to 8 cub. ft. of free air..
The weight of air concej^ned has now to be calculated.
(1) Suction air.
PxV
From equation (1) w=::;^^-;^ — — = 14:4:X
(2) Blast air.
14-7x108-75
53-2xT
53-2x521
= 8-3lb.
Weight of blast air=
8-3x8
=0-61 lb.
108-75
„ „ suctionair+blast air=8-3+0-61=8-91 lb.
The efEect of the heat released by the combustion of the fuel
is to increase the temperature of : —
(1) The suction air (8-3 lb.),
(2) The blast air (0-61 lb.),
at a constant pressure of 514-7 lb. per sq. in. abs.
The temperature of the blast air is taken to be 521° F. abs.,
and that of the adiabatically compressed air is found from
equation (6) as follows : —
Ta \Pa/ V 14-V
2-805
Therefore Tb=521 x 2-805 = 1460'
F. abs.
= 3,100° F. abs.
22 DIESEL ENGINE DESIGN
Now let H=heat added in B.T.U.
Then since the pressure remains constant,
H==-238 [8-3 (To-1460) + -61 (T„-521)]
= 2-12T„-2961
from which T,=^4^^ (7)
" 2-12
and from equation (1)
Now suppose that 0-2 lb. of fuel is added, having a calorific
value of 18,000 B.T.U. per lb., the resulting temperature will
be:—
0-2x18,000 + 2961
212
and V„ = -0064x 3100 = 19-8 cub. ft.
The expansion line CD can now be constructed in the same
way as the compression line, both being adiabatics. The value
of the terminal pressure P^ is required for calculating the
work done, and is found by means of equation (4) as follows : —
p^ [yj V108-75/ -"^^
.-. Pa = -091 X 514-7 =46-8 lb. per sq. in. abs.
Calculation of Work Done. — Referring to Fig. 8, the work
done is clearly equal to the areas BCNM plus CDRN less BARM.
Using equation (5) for the two latter, we have : —
Area BCNM = 144 x 514-7 x (19-8 -8-75) = 819,000 ft. lb.
„ CDRN = 144(514-7xl9-8-46-8xl08-74)^^g^^^^^^^^^
-41
By addition 2,619,000 „ „
BARM = 144(514-7x 8-7^14-7 xl08-75)_, ^„^ ^„^
,, ^jr 9 — 1,UZU,UU0 „ ,,
Work done by difference = 1,599,000 „ ,,
Since 1 H.P. hour = 1,980,000 ft. lb.
the fuel consumption is : —
0-2X1,980,000 „.oM. TTT-DT-
1,599.000 =-248lb.perI.H,P.hour. ,J
[>
THERMAL EFFICIENCY 23
From equation (3) the work done in compressing the blast air
is given by
914-7
8 X 14-7 X 144xloge (assuming blast pressure of
^*"' 900 lb. per sq. in.).
= 70,000 ft. lb.
so the nett work done is
1,599,000-70,000=1,529,000 ft. lb.
and the consumption of fuel per B.H.P. hour is : —
0-2x1,980,000 OKOIU T}TTT)U
=-259 lb. per B.H.P. hour.
1,529,000 ^
The mechanical efficiency being ,'„„'„_ =95-6%
1,599,000
The M.I.P. of the diagram is equal to the indicated work
divided by the stroke volume.
1,599,000 icnn/^iu i^
— ^ = 15,990 lb. per sq. ft.
100 f ^
= 111 lb. per sq. in.
This is a trifle higher than the M.I.P. usually indicated at full
load in an actual commercial engine.
The brake thermal efficiency is given by : —
^'^SQ'OQQ =0-548
0-259x18,000x778
A splendid figure for the brake thermal efficiency of an
actual engine of large size is 0-35 ; comparing this with the
above figure, it is seen that the actual engine attains 64% of
the efficiency attributed to the ideal. The indicated thermal
efficiency of the ideal engine is : —
1,980,000 =0-573
0-247x18,000x778
That of an actual engine at full power is about 0-472, the
ratio of actual to ideal being about 82%. From the above it
will be seen that so far as the thermal actions within the cylinder
are concerned, an actual Diesel Engine in good order leaves
comparatively little room for improvement as long as the
accepted cycle is adhered to. As a matter of fact, slight
deviations from the constant pressure cycle are frequently
made, and improved efficiencies obtained thereby. Most high
speed Diesels, for example, are arranged to give an indicator
24
DIESEL ENGINE DESIGN
card, showing a certain amount of explosive effect, i.e. combus-
tion at constant volume, causing the pressure at the dead
centre to rise to a figure which may be anj^hing up to about
100 lb. above the compression pressure. This is found to have
a beneficial effect on the fuel consumption, which is readily
explained on theoretical grounds ; but it is obvious that
considerations of strength must limit the extent to which this
principle is used.
Fuel Consumption at Various Loads. — If the foregoing
calculations for the fuel consumption of the ideal engine are
09
0-8
I .0-7
:i:-a
(u m
Q. u
S;-oo-4
— ™
-a
-> 0-1
n
1
\
\
,
\j
^
1
^
^Z
^
A
"<i
^
«.
Pr
»C
iJhJ -
—
Arl-uil Cngine'(ln-aicai:ea;
_
Id^at'Cnqine I'Bteke) ^
1
_
^
-ik^h^n^Jn
' 7
ra
caTe
'oT~
—
"■"
_J
10 20 30 40 50 60 TO 80 90 100 110 120
Indicated Mean Pressure lbs/ in^
Fig. 9.
repeated for various values of the quantity of fuel admitted
results will be obtained which are shown graphically in Fig. 9,
together with test results of an actual engine. The fuel con-
sumptions per I.H.P. and B.IUP. hour are plotted on a basis
of M.I.P. ; the actual engine to which the test results refer was
of the four stroke type developing 100 B.H.P. per cylinder at
full load. There are two facts to be noticed : —
(1) The fuel consumption per I.H.P. hour increases as the
M.I.P. increases.
( 2) The fuel consumption per B.H.P. hour attains a minimum
value at about 90 lb. M.I.P. in the case of the actual engine
THERMAL EFFICIENCY
25
and about 60 lb. in the case of the ideal. The reason for (1)
will be evident on comparing theoretical diagrams for various
values of the M.I. P., and the case is quite comparable to that
of a steam engine working with an early cut-off.
The point of maximum brake efficiency depends upon two
conflicting influences, viz. the indicated efficiency which
decreases and the mechanical efficiency which increases as the
M.I.P. is augmented.
It is usual in this country, when considering mechanical
efficiency, to treat the work done in driving the air compressor
as a mechanical loss, so that : —
T> TT p
Mechanical efficiency = ^ ^ _. — ".''.,
i.H.r. of mam cyunders.
Continental engineers, on the other hand, sometimes subtract
the indicated power of the compressor from that of the main
cylinders for the purposes of the above equation.
The usual British practice will be adhered to throughout
this book.
Variation of Mechanical Efficiency with Load. — Examina-
tion of a large number of Diesel Engine test results reveals the
fact that the difference between the I.H.P. and the B.H.P.
remains nearly constant as the load is varied. This fact enables
one to calculate the mechanical efficiency at any load if the full
load efficiency is known.
Example: What is the mechanical efficiency at three-quarter,
half and quarter load if that at full load is 72% ? Let the full
load I.H.P. be represented by 100, then the following tabulated
figures hold good :-
Const.
Meoh.
B.H.P.
I.H.P.
Diff.
Effioy.
Full load
. 72
100
28
•72
Three-quarter load
54
82
28
•66
Half load
. 36
64
28
•56
Quarter load .
18
46
28
•39
This method yields sufficiently accurate results for most
estimating purposes.
Mechanical Losses. — ^The work corresponding to the differ-
ence between the I.H.P. and the B.H.P. is approximately
accounted for in the two following tables, which apply to
26
DIESEL ENGINE DESIGN
medium-sized engines of good design, and of the four stroke
and two stroke types respectively : —
Four Stroke Engine — Full Load Mech. Efficy., 75%.
per cent
Brake-work 75-0
Work done on suction and exhaust strokes . . 3-0
Indicated compressor work 5-8
Compressor friction 1'2
Engine friction, opening valves, etc. . . . 15-0
Work indicated in main cyUnders , . . . 100-0
Two Stroke Engine— Full Load Mech. Efficy., 70%.
per cent
Brake- work 70-0
Indicated compressor work 6-5
Compressor friction 1-4
Indicated scavenger work 7-5
Scavenger friction 1-6
Engine friction, opening valves, etc. . . . 13-0
Work indicated in main cylinders . . . 100-0
Improvements in bearings and guides on the principle of the
well-known Michell bearing or the use of roller bearings for the
main journals and big ends suggest possibilities for reducing
engine friction which will possibly materialise in the future.
The adoption of some form of limit piston ring to prevent
e^essive pressure on the liners would also help matters in the
'same direction, besides increasing the life of the liners.
Influence of Size on Mechanical Efficiency and Fuel Con-
sumption. — Piston speed, within the range of present practice,
appears to have very little influence on either mechanical
efficiency or fuel consumption- This remark does not apply
to the abnormally low speeds obtaining, for example, with a
marine engine turning at reduced speed.
The cylinder bore is the principal factor in economy, always
assuming a reasonable ratio of bore to stroke and good design
generally.
The following table is a rough guide to the mechanical
efficiency and fuel consumptions to be expected from cylinders
of various sizes working at full load.
THERMAL EFFICIENCY
27
FoTTR Stroke
Two Stroke
Cylinder
£>iam.,
in.
Meoh.
Effcy.
%
Fuel
per
B.H.P.
hr. lb.
Fuel
per
I.H.P.
hr. lb.
Cylinder
Diam.,
in.
Meoh.
Effcy.
%
Fuel
per
B.H.P.
hr. lb.
Fuel
per
I.H.P.
hr. lb.
10
•70
•46
•320
10
•65
•51
•332
15
•73
•43
•315
15
•68
•48
•326
20
•75
•41
•308
20
•70
•46
•322
25
•76
•40
•305
25
•71
•45
•320
30
•76
•40
■305
30
•71
■45
•320
The fuel consumption per B.H.P. at loads other than full
load is readily found by first calculating the probable mechanical
efficiency, as described in the previous article, and then allow-
ing for a fall in the consumption per I.H.P. proportional to that
shown on Fig. 9 for a typical engine for the same M.I. P.
Heat Balance Sheet. — An elaborate trial of a Diesel Engine
includes the measurement of the quantity of cooling water
used, the inlet and outlet temperatures of the water and the
temperature of the exhaust gases. Apart from very slight
losses, such as radiation, etc., these data usually enable the
heat supplies by the fuel to be accounted for.
A typical heat balance is given below : —
Accounted for by indicated work .... 42%
Rejected to cooling water 34%
Rejected to exhaust 24%
Total heat supplied . . .100
A striking feature of this balance is the large amount of heat
appearing on the cooling water account, which at first sight
would appear to indicate very poor utilisation of heat within
the cylinder. It has been shown that so far from this being the
case, an actual engine in good order indicates about 80% of the
work attributable to an ideal engine.
The explanation of this lies in the fact that a large propor-
tion of the heat received by the cooling water is given out by
the exhaust gases after combustion is complete, particularly
on their passage through the cylinder cover in the case of a
four stroke engine and through the ports in the case of a two
cycle. Most of the friction work done by the piston and all the
compressor work appear on the cooling water account.
28 DIESEL ENGINE DESIGN
Efficiency of Combustion. — In all actual oil engines there is
a considerable amount of " after burning," i.e. gradual burning
during the expansion stroke. In a well-tuned Diesel Engine
this effect is not sufficiently pronounced to cause smoke even
at considerable overload, 120 lb. per sq. in. M.I.P. for example.
Exaggerated " after burning " is to be avoided as, in addition
to increasing the fuel consumption, it increases the mean
temperature of the cycle and of the exhaust stroke particularly,
and gives rise to accentuation of all the troubles which arise
from the effects of high temperature. The most prominent of
these troubles are enumerated below : —
(1) Cracking of piston crown and cylinder covers.
(2) Pitting of exhaust valves.
(3) Increased difficulty of lubricating the cylinders, resulting
in —
(a) Increased liner wear.
(b) Sticking of piston rings.
(c) Liability to seizure of piston.
(4) Increase of temperature of gudgeon pin bearing.
The above formidable list is probably not exhaustive, but is
sufficient to show the desirability of securing the best possible
conditions, apart altogether from the question of economy in
fuel and lubricating oil consumption. The attainment of good
combustion, assuming a good volumetric efficiency of cylinder
and good compression, depends more than anything upon
small points in connection with the fuel valve, which are easily
adjusted on the test bed, provided the design of the fuel valve
is satisfactory.
Entropy Diagrams. — Entropy is a convenient mathematical
concept which it is difficult, and perhaps impossible, to
define in non-mathematical terms. It is sometimes described
as that function of the state of the working fluid which remains
constant during an adiabatic precess. Entropy increases when
heat is taken in by the working fluid and decreases when it is
rejected. If heat is supplied to the working fluid at constant
temperature, then the increase of entropy is equal to the
amount of heat so supplied divided by the absolute tempera-
ture. If the temperature is variable during the process of
heat absorption, then the increase of entropy is determined by
the integration of the equation : —
THERMAL EFFICIENCY 29
, _dH where 0=Eiitropy.
^"""t" H= Heat taken in or given out.
T =Absolute temperature.
The zero of entropy may be located at any convenient level of
temperature except absolute zero. It is convenient for our
purposes to consider the entropy to be zero when
P = 14-7 lb. per sq. in. abs.
T=521°F.
The value of a diagram connecting T and (j> during the working
cycle of an internal combustion engine depends upon the
following properties : —
(1) Increasing and diminishing values of ^ denote heat
supplied and heat rejected respectively.
(2) For a complete cycle the diagram is a closed figure, the
area of which is proportional to the quantity of heat
which has been converted into work during the cycle.
The area of the diagram is given by : —
y"T d(p==/ — — — =Hi— Hj the difference between the heat
supplied and that rejected.
Construction of an Entropy Chart for Air (see Fig. 10). —
The axis of T is drawn vertically, and includes temperatures
from to 4000° F. abs. The axis of <f> is horizontal, and is
graduated from to 0-36. It will be clear that the axis of T
is an adiabatic line of zero entropy and that any selected
temperature on this line corresponds to a definite pressure,
which can be calculated by means of equation (6). Actually
it is more convenient to tabulate a series of pressures from
to, say, 7001b. per sq. in. abs., and tabulate the corresponding
temperatures. Points on the axis of T found in this manner
are the starting points of constant pressure lines.
Constant Pressure Line. P = 14'7 lb. per sq. in. — A
constant pressure line is a curve connecting T and <p when P
remains constant. It is usual to consider 1 lb. of air, and if
the specific heat at constant pressure is assumed to be 0-25
Then dH =0-25 dT
. , ,^ dH „. dT
And d<^ = -^ = -25^
Therefore = -25 log, ^— Tj being 521° F. abs.
30
DIESEL ENGINE DESIGN
Selecting various increasing values of Tj corresponding values
of fj> are calculated and plotted on the chart, a fair curve
passing through the points being the required constant pressure
line. One such line having been constructed, lines correspond-
ing to other pressures are readily drawn, since their ordin4tes
Pounds per Sq In. Absolute.
600X0400300 200150
■05 -10 -IS -20 26 -30 -35
Entropy per Lb of Cycle Air
Fig. 10.
are proportional to the temperatures indicated by their starting
points.
Constant Volume Lines. — These can be constructed in a
similar manner, using the specific heat at constant volume
instead of that of constant pressure. On Fig. 10 only one
constant volume line is shown, ^z. that corresponding to
P = 14-7 lb. per sq. in.
T=673°r. abs.
as this is required to complete the diagram by representing the
rejection of heat at constant volume at the end of the stroke.
This process is of course a scientific fiction, as the pressure in
an actual cylinder is reduced to approximately atmospheric
pressure by the discharge of exhaust gases and not by cooling
THERMAL EFFICIENCY 31
the latter. The reason for using the value 0-25 for the specific
heat must now be explained. Although the specific heat of
pure dry air is 0'238, that of the gases present in the cylinder
of a Diesel Engine is a variable quantity for the following
reasons : —
(1) The composition of the working fluid is altered by the
addition of the fuel.
(2) The specific heat of the exhaust gases increases slightly
with increase of temperature.
If these variations were rigidly taken into account the con-
struction of the entropy diagram would be a very laborious
business, and the work is greatly curtailed by adopting certain
approximations which will be described. In the first place
the specific heat is assumed to be constant and equal to the
calculated specific heat of the exhaust gases at 60° F. The
variation in the weight of the working fluid is dealt with by
first treating the diagrams as though the weight were constant
and then correcting the diagram (entropy) by increasing the
entropy and decreasing the absolute temperature of points on
the expansion line in the same proportion by which the fuel
and blast air increase the weight of the charge.
Use of the Entropy Chart. — A method of constructing the
entropy diagram, corresponding to an indicator card, will now
be described. The clearance volume of the engine must first
be ascertained and the card accurately cahbrated. Points on
the indicator diagram are then marked, corresponding to every
15 or 30° or other convenient division of revolution of the
crank past the top dead centre. Each of the 12 or more points
so marked is given a reference number and the absolute
pressure in lb. per sq. in., and also the volume in any arbitrary
units (hundredths of an inch on the diagram, for instance)
corresponding to each point is read off and tabulated. The
■ apparent temperature corresponding to each point is then
calculated by means of equation (1) on the assumption that the
initial temperature just before compression is, say, 673° P. abs.
Points on the entropy diagram corresponding to the selected
points on the indicator diagram are now found by following
the appropriate constant pressure lines until the calculated
temperatures are reached, thus obtaining the apparent entropy
diagram, which requires to be corrected in accordance with
the preceding article.
32
DIESEL ENGINE DESIGN
The entropy diagram shown in Fig. 10 has been constructed
in the manner described, and the data are given below : —
M.I.P. shown by indicator diagram, 82 lb. per sq. in. (see
Pig. 11).
ISO
135
ISO
M.I
P =
105 '90 75
82 lb s. per
GO
S(|. n.
4! 30
3
u
';3
500
300 fe-
Deqrees
anerT.D.C.
200 -|
100.1
Atmospheric Line
Fig. 11.
Initial temperature of suction air before compression
assumed to be 673° F. abs.
Absolute temperature (apparent) calculated from PV=kT
where P=Pressure in lb. per sq. in. abs. (scaled off diagram)
V= Volume measured in linear inches off diagram
T=Temperature in degrees F. abs.
Initial conditions are P=14-7, V=6-38, T=673.
Whence k = -1395.
Figures are given in the table below : —
Ref. No.
Degree after
firing centre.
V. in
in.
0-48
P. in
lb. per sq. in.
Apparent
Temperature.
1
490
1690
2
15
0-63
498
2260
3
30
0-95
421
2880
4
45
1-53
251
2760
5
60
2-17
163
2540
6
75
2-94
114
2400
7
90
3-Z5
4*6
84
2260
8
105
65
2130
13
180
6-38
14-7
673
19
265
2-94
44
930
20
300
2-17
61
950
21
315
1-53
99
1090
22
330
0-95
184
1250
23
345
0-63
331
1500
THERMAL EFFICIENCY 33
The apparent entropy diagram was plotted from the above
values of P and T, and the corrected diagram, constructed in
accordance with the preceding article, is shown in Fig. 10.
Area of corrected entropy diagram 9-90 sq. in.
Temperature scale . . . 1 in. =500°.
Entropy scale .... lin. = -05°.
Therefore work done per lb. of suction air is given by : —
9-90 X 500 x0-05 = 248 B.T.U.
1 lb. of suction air @ 673° F. abs. occupies 16-9 cub. ft. Since
clearance volume = 8% of stroke volume, therefore corre-
sponding stroke volume=;16-9-^l-08 = 15'65 cub. ft. There-
fore work done by suction air, according to the indicator card,
@ 82 lb. per sq. in. M.I.P. is given by : —
82xl44xl5-65 ^^33^^^
778
This figure is about 4% less than that shown by the entropy
chart, and suggests that perhaps the assumed temperature of
the suction air before compression (viz. 673° F. abs.) is too
high.
Note. — ^The low pressures at the beginning of the compression
stroke and the end of the expansion stroke have not been used
in the construction of the entropy chart for the following
reasons : —
(1) Low pressures are very difficult to scale off the indicator
card with any degree of accuracy.
(2) The low pressures indicated on the card are invariably
erroneous, unless the greatest care has been exercised
in taking the card, and that with a suitable indicator
in perfect order. Many of the reputable makes of
indicator, though quite suitable for steam engine work,
give very inaccurate results when applied to internal
combustion engines.
The following points should be noticed : —
(1) The compression line deviates from the true adiabatic in
a manner which indicates that heat has been lost to the
cylinder walls.
(2) The expansion does not become adiabatic until well after
half stroke, showing that there is a certain amount of
" after burning."
34 DIESEL ENGINE DESIGN
Area of Entropy Diagram. — If the diagram has been care-
fully done the area of the diagram in heat units should corre-
spond fairly closely to the work shown on the indicator diagram.
Deviations from equality may be due to : —
(1) Variation in the specific heat at high temperatures.
(2) Incorrect assumption of the initial temperature of the
induced air.
Owing to the approximations referred to above there is
generally a discrepancy of a small percentage between the
amounts of work shown by the indicator card and the entropy
diagram. Also the total area under the upper boundary of the
entropy diagram should correspond to the total heat supplied
per lb. of suction air less the heat discarded to the water-jacket
on the expansion stroke. Investigation of the sort described
seems to indicate that not more than about 10% to 15% of
the total heat supplied by the fuel is lost in this way at
100 lb. M.I.P.
Specific Heat of Exhaust Gases. — The specific heat of a
mixture of gases is the sum of the products of the specific heats
of the constituents into the proportion by weight in which the
constituents are present.
The specific heats (at constant pressure) of the constituent
gases present in the exhaust of an oil engine are given below : —
Water vapour . . '. 0-480 (Varies considerably with
Nitrogen
Oxygen
CO2 . .
CO
0'247 temperature)
0-217
0-210 (Varies considerably with
0-240 temperature)
The average composition of air and the specific heat derived
from that of its constituents are given below : —
N2 . . 75-7%x0-247 = -187
O2 . . |2-7 x0-217 = -049
H2O . . J^ X 0-480:=^ -007
Total . _100 -2^
This is about 3% higher than the accepted value for pure
dry air.
The approximate composition of exhaust gases assuming
complete combustion is readily calculated as foUows : —
Data M.I.P. - 100 lb. per sq. in.
THERMAL EFFICIENCY 35
Fuel per I.H.P. hour -0-31 lb.
r Carbon, 86% |
Composition of fuel . I Hydrogen, 13% r by weight,
(assumed). I Oxygen, 1% J
One H.P. hour = 1,980,000 ft. lb.
.". Volume swept by piston per I.H.P. hour=-^— — ^— -
^ -^ ^ ^ 100x144
= 137-5 cub. ft.
Clearance volume, say, 8%= 11 -0 ,, ,,
Total weight of suction air =148-5 (assuming perfect scavenge
of clearance space).
w - i-j. t 4.- ■ 148-5 X 14-7x144 onoiu
Weight of suction air=^ -— - — — -r =8-78 lb.
@ 212° F.
Free volume of blast air @ 8% of stroke volume @ 60° F.
= •08x137-5 = 11 cub.
Weight of blast air=i^^^^^^^=0-83 lb.
.: Total air = 8-78+0-83=9-611b.= 97%
Weight of fuel . =0-31 „ = 3%
Total mixture . 9-92 „ 100%
Composition of mixture before combustion is given by : —
(Na. . . -76x97=73-7%
Air O2(22-7x97) + (-01x3)=22-0%
(H^O . . -015x97= 1-5%
*"^Hh2. . . -13X3= 0-4%
100-1%
Combustion of carbon and hydrogen takes place in the pro-
portions given by : —
C+0j=C02
12+32=44 by weight
And H2+0=H20
2 + 16 = 18 by weight
36 DIESEL ENGINE DESIGN
The following table shows the composition after combustion
and the specific heat (constant pressure) of the mixture : —
Per cent. Specific heat. Product.
Na 73-7 X -247 = -1820
CO2 2-5x44^12 . . 9-2 X -217 = -0191
O2 22-(32x2-5-i-12) . 12-1 x -210 = -0254
-(16x0-4^2)
H2O l-5-(18x0-4-2) . _5J[ X -480 = - 0245
1004
Specific heat of exhaust gases = '2510
Repeating the above calculation for different values of the
M.I.P., the following figures are obtained : —
M.I.P. lbs. per sq. in. 30 45 75 105 130 155
Specific heat . . -245 -247 -248 -260 -252 -254 -256
Literature. — For thermodynamics of the internal combus-
tion engine consult : —
Wimperis, H. E., " The Internal Combustion Engine."
Judge, A. W., " High Speed Internal Combustion Engines."
Clerk, Sir Dugald, "The Thermodynamics of Gas, Petrol,
and Oil Engines."
For test results see : —
Chalkley, A. P., " The Diesel Engine."
Dalby, W. E., "Trials of a Small Diesel Engine."— Inst.
N. A., April 2nd, 1914.
WUkins, F. T., "Diesel Engine Trials."— Inst. M. E.,
October 20th, 1916.
— " Description and Tests of 600 B.H.P. 2 stroke cycle
direct reversible Diesel Marine Engine." — Engineering, Dec.
22nd, 1916,
CHAPTER III
EXHAUST, SUCTION AND SCAVENGE
Renewal of the Charge. — In the theoretical study of heat
engines the charge of working fluid is supposed to remain
enclosed in a working cylinder and to undergo a cycle of
physical changes due to the introduction of heat from, and
discharge of heat to, external sources by conduction through
the walls. In actual engines of the internal combustion type
one constituent of the charge, viz. the oxygen, takes an active
part in the chemical processes which constitute the source of
energy. In such engines, therefore, the air charge must be
renewed periodically. In all existing types of oil engine the
charge is renewed as completely as possible for each cycle of
thermal changes.
The way in which this is done in four stroke and two stroke
Diesel Engines has been described in general terms in Chapter I.
In the present chapter it is proposed to discuss the questions
of the discharge of exhaust gases and the introduction of a
new air charge from the quantitative point of view, apart from
the consideration of the mechanical details, such as valves,
cams, etc., of which these processes involve the use.
At the outset it will be necessary to state in a form con-
venient for application, the laws governing the flow of gases
through orifices.
Flow of Gases through Orifices. — Fig. 12 is intended to
represent a chamber containing gas at a pressure and tempera-
ture maintained constant at the values P^ and Tj, and from
which a gaseous stream is issuing through an orifice into the
surrounding space, which is filled with gas at a constant pressure
Pa-
In the first instance, it will be supposed that P^ is very much
greater than Pj. Then, according to the elementary theory of
the flow of gases, it can be shewn that if the gas composing the
stream expands adiabatically and all the work done is expended
37
38
DIESEL ENGINE DESIGN
in increasing the kinetic energy of the stream, then the velocity
of the latter will increase as expansion proceeds and attain a
maximum value given by : —
V = V2g J K„ (Ti-Ta) = 109-6V(T7=T;) (1)
where : —
V=maximum velocity of stream in ft. /sec.
g =32-2 ft. /sec. 2 (acceleration due to gravity).
J = Joules' equivalent (778 ft. lb. /B.T.U.).
Kp=Specific heat at constant pressure (for air Kp = -238
B.T.U. /lb. deg. F.
T2=Temperature attained by the stream after adiabatic
expansion from P^ to Pj, and is given (for air) by : —
(p \ 0.285
In equation (1) the velocity in the chamber is supposed to be
negligibly small compared with V. For values of Pi up to
P, ,T,
Ps
Fig. 12.
70 X 144 lb. per sq. ft., Pa being atmospheric, equation (1) has
been f oimd to agree with experiment within two or three per
cent, which is ample accm'acy for our purpose.
It may, however, be mentioned in passing that experiments
at high pressures, and particularly with steam, indicate that
the elementary theory on whreh equation (1) is based stands
in need of correction.
It is to be observed that equation (1) is equally valid for any
stage of the expansion of the stream, so that if any pressure
value P^ be selected lying between Pj and Pj, then the
velocity at this stage wiU be given by equation (1) with P^
substituted for Pg. If values of the velo6ity V^ be calculated
for various values of P^ and the specific volumes v^ corre-
EXHAUST, SUCTION AND SCAVENGE 39
spending to these values, then the ratios ==^ will be proportional
to the areas of the stream at the several stages of expansion.
On making such a calculation it wiU be found that the area of
the stream at first contracts and afterwards converges to a
final value corresponding to the maximum velocity. In other
words, the stream has a neck or throat, as shown in Fig. 12.
The pressure Pq at the throat is known as the critical pressure,
and for air is equal to O-SS P^.
Furthermore, it is evident that if Pa is equal to Pc (a con-
tingency which was ruled out at the beginning of the discussion),
then the stream will converge to its throat area and remain
parallel instead of diverging.
Supposing again that P2<Pc, it is evident that for a given
size of orifice the discharge will be a maximum if the throat of
the stream occurs at the orifice, and in this case the dis-
charge in lb. per sec. wiU be given by : —
Q=AVoWc (3)
where Q=discharge in lb. per sec.
A=area of orifice— ft. ^
Vc=throat velocity— ft. /sec.
W£,= weight in lb. per cub. ft. of the air at the conditions
of temperature and pressure obtaining at the
throat.
So that the discharge is independent of the back pressure Pa
so long as P2<Pc.
On the other hand, there is no guarantee that the throat of
the stream will coincide with the orifice in every case, so that
in general the discharge given by equation (3) has to be multi-
plied by a discharge coefficient less than unity, in order to give
the discharge observed by experiment.
For air and exhaust gases the following figures may be
used : —
Discharge coefficient for sharp-edged orifices or ports— 0-65
„ „ ,, mushroom valves . . .0-70
The velocity calculated from equation (1), multiplied by the
appropriate coefficient, may conveniently be called the
apparent velocity referred to the actual area of the orifice.
The Suction Stroke. — By way of application of the preceding
formulae, we may consider the suction stroke of a Diesel Engine.
The retreating piston creates a partial vacuum and air passes
40
DIESEL ENGINE DESIGN
in from the atmosphere via the inlet valve in consequence.
Supposing it is desired to limit the pressure difference to 1 lb.
per sq. in., then —
Pi = 14-7xl44 P2 = 13-7xl44 Ti=say 520° F. abs.
'13. 7\ 0.285
Then T, = 520
14-7
= 509-7° F. abs.
And V = 109-5v'520-509-7 = 351 ft. /sec.
Using a coeflficient of 0-7 for the value, the apparent velocity
referred to the valve area is
351x0-7 = 246 ft. /sec.
o7?00
V
V)
^000
u.
$800
u
_o
^600
.ij
0400
t-
200
^
^
-^
^
^
/
X'
/
/
/
14-7 13-7 12-7 11-7 10-7 97 8-7 7-7 6-7
Suction Pressure Lbs. IN^ Absolute
Fio. 13.
5-7
^ If the pressure difference is to be constant at 1 lb. /in.^ then
the valve and its operating cam must be so designed that
Instan taneous piston speed (ft. /sec.) = Piston area
246 » Instantaneous valve area.
As indicated in Chapter I, it is better to allow the inlet valve
to open before the inner dead centre, and in order to obtain a
maximum charge of air the valve should not seat until 20° or
30° after the outer dead centre has been passed.
Fig. 13, which is a curve connecting calculated velocity and
suction pressure, is the result of repeating the above for
different values of P 2 on the assumption that
Pj = 14-7 X 144 and Ti=520° F. abs.
EXHAUST, SUCTION AND SCAVENGE
41
Scavenging of Two Stroke Cylinders. — Before the scavenge
air supply is put into communication with the cyUnder,
whether by valves or ports, the exhaust slots should have been
sufficiently uncovered for the cyUnder pressure to have fallen
to practically atmospheric pressure. As wiU be shown later,
this takes place with great rapidity owing to the high tempera-
ture of the gases. On this account the introduction of scavenge
air may be assumed to take place against a pressure differing
but very slightly from atmospheric. This, however, does not
apply to the later stage of the process in those valve scavenged
or controlled port scavenge engines in which scavenge air is
WOO
3ack Pressure 15 Lbs
./In?
___ — •
t/Sec
,^-^
-"^
>£00
u
^
o
^400
y
/
/
rhroa
/
/
(
lb 16 17 18 19 20 21 22 23
Scavenge Pressure Lbs/In? Abs-
Fig. 14.
2a
forced into the cylinder after the exhaust ports have been
covered. During this latter stage the cyhnder pressure con-
tinually increases until either the scavenge valves (or ports)
close and compression begins, or until the cyhnder contents
and the scavenge air supply are in equilibrium.
The first stage of the process may be dealt with in a similar
manner to that indicated in the preceding article, and Fig. 14
shows the calculated throat velocity of the scavenge air plotted
against the absolute scavenge pressure, on the assumption that
the back pressure is 15 lb. per sq. in. abs. and the temperature
of the scavenge air supply is 130° F. (591° F. abs.), this being
a good average figure found in practice.
Before applying the above to a concrete example some
observations will be made on port and valve areas.
If a fluid flows for a time " t " through an orifice of constant
42
DIESEL ENGINE DESIGN
area " A," with a velocity " V," then the volume " v "
discharged is evidently given by : —
v=V(Axt)
If on the other hand A is variable, then
v=V/A-dt
The quantity yA-dt known as the time integral of the area
is readily found by plotting A as ordinate against " t " as
abscissae, as in Fig. 15, which exhibits the opening area of a
valve from the time it lifts (t=o) to the time it seats.
The time integral of the valve area for the whole interval is
the area under the curve. For any other interval from, say,
t=ti to t=t2the time integral of the area is the area bounded
by curve, the axis of t and the two ordinates which define
Fig. 15.
tj and tg. This area is shown shaded on the diagram for
particular values of t^ and tg.
It is often convenient to take as the unit for " t " i^th sec.
or xrnyrtli sec, or even 1 degree of revolution of the crank-shaft.
Also it is very convenient to plot on a " t " base a curve
whose ordinates representyA-dt from t=o.
This has been done on Fig. 15. It will be noticed that
between the instants t=ti sAd t=t2 the value ofyA-dt has
increased by an amount X. The volume discharged between
these instants would therefore be (X-V).
Example of Scavenge Calculation .-j-Data for two stroke
engine : —
Bore of cylinder 10"
Stroke 15"
Revolution per minute . . . 300
EXHAUST, SUCTION AND SCAVENGE
43
Two scavenge valves 3 J" diam., maximum lift 1 ", opening 25°
before bottom dead centre and closing 60° after. Lift curve
harmonic, i.e. the lift plotted on a time base is a sine curve.
Exhaust ports occupy 60% of the circumference of the cylinder
bore and become uncovered by the piston 15% before the end
of the stroke. The connecting rod is 5 cranks long. Fig. 16
has been drawn to show the relation between percentage of
Fig. 16.
stroke and the number of degrees between the crank position
and the bottom dead centre. From this diagram it will be seen
that the ports are uncovered 50° before B.D.C. and covered
again 50° after B.D.C. Compression space 7% of the stroke
volume. Free air capacity of scavenger 50% in excess of stroke
volume of impulse cylinders.
„ , , -785x102x15
Stroke volume =-
1728
Compression space =0-07 x 0-68
Stroke volume + compression space
Volume of scavenge air (at atmospheric
pressure and temperature) deUvered per
cylinder per revolution=0'68 x 1'5
Maximum exhaust port area
_ 0-6XTrxl0x0-15xl5
" 144
Maximum scavenge valve area
_ 2X7rX3-5xl
~ 144
=0-68 ft.3
=0-05 „
=0-73
= 1-02
=0-295
= 0-153 ft. 2
44
DIESEL ENGINE DESIGN
Intermediate values of the exhaust port and scavenge valve
areas have been plotted on a crank angle base on Fig. 17.
Number of seconds corresponding to 1 degree of revolution
of the crank-shaft
In Fig. 17 values of/A-dt for both exhaust and scavenge
areas have been plotted in accordance with the preceding
article. In particular, since the scavenge valve lift is harmonic,
Degrees Before B.D.C. Degrees AFter B.D.C.
Fig. 17.
the mean area is one-half the maximum, and the final value of
fA-dt is given by : —
^^x0-000556x85° = 10-3x3-61 ft.-'secs.
And since the port area curve is nearly parabolic the
maximum value ofy A-dt for the ports is given by : —
0-295 Xfx0-000556xl00° = 10-3x 10-94 ft.3 sees.
The intermediate values are readily found by planimetering
the areas under the port and^alve area curves.
Scavenge Air Pressure. — ^A first approximation to the value
of the pressure in the scavenger air pipe required by the condi-
tions of the problem is readily obtained, on the assumption
that the scavenge air is delivered against a constant pressure
of 15 lb. per sq. in. abs.
Volume discharged . . .=l-02ft.^
/A-dt =3-61x10-3
EXHAUST, SUCTION AND SCAVENGE 45
1-02 X 10"*
Apparent velocity . . .=■ — — — =283 ft. /sec.
Dividing by a discharge coefficient of 0-7 the
" calculated " velocity is 283 ^0-7 =405 ft. /sec.
and the corresponding scavenge air pressure from Fig. 14
is : —
= 16-35 lb. perin.^abs.
This figure cannot, however, be accepted as final, since the
effect of discharging the excess of air through the exhaust ports
has been ignored.
Assume for trial a scavenge pressure of 17 lb. per in.^ abs.
The calculated velocity from Fig. 14 is 498 ft. /sec. and the
apparent velocity
=498x0-7 = 349 ft. /sec.
The value oty^A-dt which must be attained to fill the stroke
volume and clearance space (0-73 ft.*) is therefore : —
^3=2.09X10-
This corresponds to point A on Fig. 17, and shows that with
the assumed scavenge pressure the cylinder would be filled
with scavenge air at about 21° after the bottom dead centre.
From this point until the point where the exhaust ports are
covered three processes are occurring simultaneously, viz. : —
(1) Scavenge air is escaping out of the exhaust ports.
(2) Scavenge air is entering the cylinder and raising the
pressure of its contents.
(3) The piston is rising and reducing the volume of the
cylinder contents, at the same time tending to raise
the pressure.
The pressure which exists in the cyUnder at the instant at
which the piston covers the ports may conveniently be termed
the " initial charge pressure," and will be denoted by Pj.
Now if the value of Pj were equal to the scavenge pressure,
which is its upper limit, the amount of air introduced into the
cylinder in the interval under consideration would be about
equal to : —
f/A-dt (from 21° to 50° after B.D.C.) X 349
(for valves)
46 DIESEL ENGINE DESIGN
From Fig. IT/A-dt for this interval is l•4:XlO"•^ so the
vdlume introduced is
f X 1-4x10-3 X 349=0-326 ft. 3
Adding to this the volume already introduced, viz. 0-73 tt.^,
we obtain 1-056 ft. 3, which is a trifle in excess of the required
quantity.
The value assumed for the scavenge pressure, viz. 17 lb. /in.^
abs., is, therefore, probably not far out.
The precise value of P, is a matter of considerable im-
portance, as on it depends the value of the charge weight, and,
consequently, the power capacity of the cylinder. Its pre-
determination, however, appears to be a difficult matter, and
in practice it is usually adjusted experimentally by advancing
or retarding the scavenge cam, according as the value of Pj
found by light spring indicator diagram is too low or too high.
In the former case an adjustment is easily effected by putting
a baffling diaphragm in- the exhaust pipe.
The above calculations, however, are a sufficient check on
the design to ensure the provision of suitable valve or port
areas.
The volume (@ 15 lb. /in. ^ abs.) of scavenge air lost through
the exhaust ports is roughly equal to
^/A-dt (from 21° to 50° after B.D.C.) X 498 x 0-65
(for ports)
From Fig. 17/A-dt for this interval is 2-15 x 10" ^ ft.« sec,
so the volume required is
^X 2-15x10-3x498x0-65=0-23 ft. 3
So that the air charge is equivalent to (1-02— 0-23) =0-79 ft.*
@ 15 lb. per sq. in. abs.
Its actual volume at the point when the piston covers the
ports is 0-05-)-(0-85x0-68)=0-63 ffc.»
and its pressure is therefore about
15x0-» ,^ ^,, ,. „
-^;g3-=18.8lb./in.^
This figure being 1-8 lb. /in.^ above the assumed scavenge
pressure of 17 lb. /in.^ indicates that the latter figure is in-
sufficient for the supply of the requisite quantity of air under
the conditions specified. The true value probably lies some-
where between the two.
Exhaust of Two Stroke Engines. — ^As already mentioned,
EXHAUST, SUCTION AND SCAVENGE 47
it is essential that the scavenge receiver should not be put into
communication with the cylinder until the contents of the
latter have fallen to a pressure almost equal to that of the
scavenge air, by the release of the products of combustion,
through the exhaust slots. This process usually takes a period
of time equivalent to 20 to 30 degrees of revolution of the
crank-shaft. The calculation of this period is much facilitated
by the fact that in the interval considered the port area is
increasing very approximately in proportion to the time (see
Fig. 17). It can easily be shown that if during an interval from
t=o to t=ti an orifice area increases uniformly with respect to
time from O to A^ and the velocity of efHux also varies uni-
formly with respect to time from V, to Vj, then the discharge
is given by : —
Q=^Mv,+iV„) — (1)
If A represents ft. 2, t sees., and V ft. /sees, of an incom-
pressible fluid, then Q represents cub. ft.
If, as in the case we are about to consider, V represents
lb. per sec. per unit of area, then Q represents lb. We proceed
to apply equation (1) to the exhaust period of the two stroke
engine specified in the previous article.
Exhaust Calculation. — Data : Cyhnder pressure at the point
at which the ports become uncovered, 60 lb. /in. * abs. Temper-
ature at the same point is equal to
initial charge temperature x 60_ 670x60
— say r-;
,, ,, pressure. 18
=2240° F. abs.
Pressure in exhaust pipe, 15 lb. /in.* abs.
Problem : To find the position of the crank when the
cylinder pressure has fallen to 18 lb. /in.* abs./
Assume that the charge has the thermal properties of pure
air and that the expansion is adiabatic.
At the instant when the ports first become open the volume
of the charge is : —
0-05+0-85x0-68=0-63ft.3
and the weight of the charge is : —
(60 x 144 x 0-63) -^(63-2 x 2240) =0-0456 lb.
When the cylinder pressure has fallen to 18 lb. /in.* abs. the
48 DIESEL ENGINE DESIGN
volume is not certainly known, but wiU not differ much from
0-05+0-92 x0-68=0-675 ft.^ and its temperature will be :—
2240(4^1 =1590°F. abs.
/18^Y28B_
\60-0/ ■"
and the charge weight is reduced to
(18 X 144 xO-675) ^(53-2 x 1590)=0-0207 lb.
The weight discharged in the interval is therefore : —
0-0456 -0-0207 =0-0249 lb.
The next step is to calculate the rate of discharge per unit
of port area.
At the higher pressure of 60 Ib./in.^ the throat pressure is
0-53 x 60=31-8 lb. /in. 2 abs., and the throat temperature is : —
(Ql .C\-288
^j =1870° F. abs.
The calculated throat velocity is therefore : —
109-5 V2240-1870=2110 ft. /sec.
and the apparent velocity =2100 x 0-65 = 1365 ft. /sec.
Now the specific volume at the throat is : —
53-2x1870 „,„,,, 1, •
^j-:g-^=21-8ft.3perlb.
The rate of discharge is therefore : —
— — =62*7 lb. per sec. per ft.* of port area.
21-8
At the lower pressure of 18 lb. /in.'' the throat pressure is
15 lb. /in. 2 abs., and the throat temperature is : —
2240(^y285 = 1510° F. abs.
The calculated throat velocity is therefore : —
109-5^1590-1510 = 978 ft. /sec.
and the apparent velocity =§78 xO-65=635 ft. /sec.
Now the specific volume at the throat is : —
53-2X1510 or, OX^ lU
— - — — — =37-2 ft. 3 per lb.
15x144 ^
and the rate of discharge is therefore : —
635
— — =17-1 lb. /sec. per ft.* of port area.
EXHAUST, SUCTION AND SCAVENGE 49
We now make the assumption that the rate of discharge
per unit area changes its value from the higher to the lower
value uniformly with respect to time, so that equation (1)
can be applied. We then have : —
0-0249=
3
from which Aj t^ = 10" ^ X 1 -54
and7'A-dt = i Aj ti (approx.) = 10-3x0-77 ft.^ sec.
Reference to Fig. 17 shows that this value occurs about
18 degrees after the ports begin to open.
Alternative Method of Calculation. —
Let :—
t=time in seconds, counting from the instant when the
ports begin to be uncovered by the piston.
P=pressure of cylinder contents in Ib./ft.^ during the
period considered, so that P varies from 60 x 144 to
18x144.
T == temperature of cylinder contents deg. P. abs. (variable).
w= weight in lb. of cylinder contents (variable).
A=port area in ft.^ (variable).
v= volume in ft.^ of cylinder contents (actually varies
during the period considered, but treated as constant)
=0-65 ft. 3
Vg= specific volume in ft.* /lb. at the throat of the stream
issuing from the ports.
V=throat velocity in ft. /sec.
dP
f (P) =--T — hA=the rate of pressure drop per unit of port area,
dtj
then^=f(P)xA
dt
and/A- dt=/j^
f (P) has now to be calculated.
KQ.O
P.v = 53-2wT .-. P- w.T.
V
, dP 53-2 /^dw , dT\
and -r--= ( T.-rr+W.- *
dt V \ "dt 'dt/
53-2 /^ dw , dT dP\
T-dt+^-dp-d^)
50 DIESEL ENGINE DESIGN
so that
dP
53;2 / dw\ 53-2 dw
V V 'dt / V dt
/, 53-2 dT\ / p dT\
(1-— •^•dp) (l-fdP/'
AT dw . V
Now -r- =A.—
dt V.
/|J\U. ZIJ5
and since T=To f.^) , where T^ and P^, are the initial
^^"^ values of T and P,
dT 0-285 T„P«'- 285-1)
dP P^o-^ss
P dT
and =--^=0-285 (i.e. constant),
dP 53-2 V
so that f(P)=^-^ A = V" V
^ ' dt V v„
0-715
Values of f (P) are easily calculated for evenly spaced values
f dP
of P and the integration of \-TTf^, can be effected by Simpson's
rule. J *(^)
The table opposite shows the application of this method
to the example of the preceding article.
Po=601b./in.2abs. To=2240° F. abs. V=0-65ft.3
For convenience P has been expressed in lb. /in.^, instead of
lb. /ft. 2, so that a factor of 144 is required. Since the interval
between successive values of P is 10-5 lb. /in.", we have : —
r dP
/A-dt=^ =10-5x144x96-18x10-8
f(P) 3
=0-485x10-3
Dividing by the discharge coefficient 0-65 the required value
of^'A-dt is :— •
0-485x10-3 „ ,. -^ „
— — =0-74 ft.^ sec. H- 1000
0-65
which agrees with the value obtained previously within about
4%-
The value of the discharge coefficient (0-65) has been found
by comparing the results of calculations similar to the above,
with the information afforded by light spring indicator cards.
EXHAUST, SUCTION AND SCAVENGE 51
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52 DIESEL ENGINE DESIGN
It is also in good agreement with experiments on the flow of
gases through sharp-edged orifices. It is by no means easy,
however, to obtain light spring indicator diagrams which are
at all reliable with an ordinary indicator. The fall of pressure
is so rapid that the shape of the card is distorted by very little
indicator stickiness, and oscillations are almost inevitable.
With care the latter may be approximately allowed for,
but cards \ showing appreciable indicator stickiness must be
rejected. These troubles may be largely eliminated by using
an optical indicator, which is perhaps the only instrument
well adapted to this class of investigation.
Fig. 18 shows a light spring diagram taken from a two stroke
Diesel Engine with an ordinary indicator in good order.
Fig. 18.
Literature. — ^Funck, G., "Two Stroke Engines with special
reference to the Design and Calculation of Ports." — Auto-
mobile, Engineering, May, 1918, et seq.
Fetter, H., "The Escape of Exhaust Gas in Two Stroke
Engines." — Engineering, January 4th, 1918.
Thompson Clarke Research, "Air Flow through Poppet
Valves." — Report to U.S. Nat. Advisory Committee for
Aeronautics. See Engineering^ January 10th, 1919.
CHAPTER IV
THE PRINCIPLE OF SIMILITUDE
The properties of similar structures, under equivalent condi-
tions, have been known for a long time, and are implied in most
of the empirical formulae used in machine design. At the
same time, there is a great deal of misconception about the
matter in the minds of many to whom a correct appreciation
of the principles of similitude would be of considerable value.
Definition of Similar Engines. — For the purposes of this
discussion, two or more engines are said to be similar when the
linear dimensions of every part of one engine bear a constant
ratio to the linear dimensions of the corresponding parts of the
other engine or engines. Stating the same condition in another
way, any dimension of any part (the diameter of the gudgeon
pin, for example) can be expressed as a fraction or proportion
of some other dimension (the cyhnder bore, for example), this
fraction or proportion being constant for all similar engines.
It follows from the above that two engines to be similar
must have the same bore to stroke ratio.
In practice there are a considerable number of deviations
from strict similarity between different sized engines of the
same type, built by the same maker, a few being noted
below : —
1 . The bore or stroke ratio is subject to variation.
2. Thicknesses of metal generally bear a larger fractional
ratio to the cylinder bore in small engines than in larger
engines.
3. Studs, bolts and other small gear are made relatively
heavier in the smaller sizes, to avoid damage by careless
handling, etc.
Equivalent Conditions. — Similar engines may be said to be
under "equivalent conditions" when the piston speed is the
same and the indicator card identical for both engines.
63
54 DIESEL ENGINE DESIGN
In the past it has been customary to reserve the higher
piston speeds for the larger sizes of engines; the modern
tendency, on the other hand, is to use the same piston speed for
all sizes. The former practice appears to have secured uniform
durability as measured by the useful life of the engine for all
sizes, whereas the latter undoubtedly results in the smaller
engines being subject to more rapid depreciation than the
larger. Without appreciable error the indicator cards, so far
as they influence the stresses in the various component parts
of the engine, may be taken as identical for all sizes, though it
is worth mentioning here that it is found advisable in practice
to work with a smaller M.I.P. in the case of large cylinders.
Properties of Similar Engines. — ^To avoid repetition equi-
valent conditions wiU be impUed when use is made of the term
"similar engines," unless the contrary is specified. Since all
parts of similar engines are in proportion their relative size may
be expressed by any linear dimension of any part, and for this
purpose it is very convenient to select the cylinder bore, as
this is the dimension which, assuming constant piston speed,
determines the power of the cylinder. Consider a whole series
of similar engines having cyhnders of different bores. Now the
piston load at any point of the stroke will be proportional to
the bore^, since the indicator card is the same for all. Now
the area of the main bearings wiU also be proportional to the
bore* since both linear dimensions contributing to the bear-
ing area are proportional to the bore. From this it follows
that the bearing pressures are the same for all the engines.
Without detaihng the matter further, it suffices to state that
with similar engines aU the bearing pressures and stresses
(including inertia stresses )^ of corresponding parts are identical.
Dealing in the same way with rubbing speeds of bearings,
velocities of gases, etc., it is found that these also are identical.
From these facts it follows that given a satisfactory engine
other engines made from the^ame designs but to a different
scale (within rational limits) will also be workable machines,
though for practical reasons such a procedure carried out in
minutiae would not be desirable. Actually, the foremost
makers of Diesel Engines have carried out the principle of
similarity remarkably closely.
Relative Weight of similar Engines. — ^Weight being propor-
tional to volume (assuming the same materials are always used
for corresponding parts), the weights of similar engines are
THE PRINCIPLE OF SIMILITUDE 55
proportional to the bore *. Hence in comparing two different
constructions of engine the weight per cubic inch of bore ^ is a
convenient criterion of the heaviness of construction. Figures
for actual engines will be given later. Again, since the power
varies as the bore^ the weight per horse-power varies as the
bore with similar engines. Owing to departures from strict
similarity, some of which have been mentioned above, the
weight per horse-power does not in practice increase propor-
tionally as the bore is increased. The practice of adopting
higher piston speeds for the larger size of engine tends in the
same direction, with the result that the weight per B.H.P.
shews comparatively small variation over a large range of
powers.
Circumstances which act in the reverse direction are the
necessity for crossheads and guides and slight reduction of
M.I.P. in the larger sizes.
In comparing the weight per B.H.P. of Land Diesel Engines
the number of cylinders must be taken into consideration for
two reasons : —
1. The cam-shaft driving gear is usually the same for a six
or four cylinder engine as for a single cylinder engine
of the same bore and type, and consequently bears a
larger proportion to the total weight in the case of the
engine with the smaller number of cyUnders. The
same applies to most of the accessories, such as starting
bottles, fuel filters, etc., if these are included in the
weight.
2. The weight of the fly-wheel is generally less for a three
cylinder engine than for a single cylinder, the bore
being the same, though it will be shewn later that
further reduction of fly-wheel weight is not advisable
for four and six cylinders.
A summary of the properties of theoretically similar engines
is given below.
Piston speed constant. M.I.P. constant.
1. Linear dimensions proportional to the bore.
2. Revolutions per minute inversely as the bore.
3. Rubbing velocities and gas velocities are the same.
4. Bearing pressures are the same.
5. Stresses are the same (including the inertia stresses).
56 DIESEL ENGINE DESIGN
6. Elastic deflections proportional to the bore.
7. Natural frequencies of vibration proportional to the
R.P.M.
8. Loads (pressure and inertia) proportional to the bore 2.
9. Horse-power proportional to the bore^.
10. Weight proportional to the bore*.
11. Weight per B.H.P. proportional to the bore.
12. Weight per unit of bore ^ the same.
The above considerations justify to a great extent the
common practice of comparing different designs by expressing
Fig. 19. Four Strokb Land Engine.
the dimensions of corresponding parts as fractions of the bore
regardless of whether machines compared are of the same
approximate size or not. The remainder of this chapter will
be devoted to applications.
Weight of Diesel Engines. — From the purchaser's point of
view the weight of a power installation per B.H.P. is a matter
of some importance, and Fig. 20 shews how this quantity varies
with different sizes of four stroke land engines of the weU-
known " A " frame type. It will be observed that the weight
per B.H.P. increases with the size of the cylinder, but not as
rapidly as the principle of simiHtude would lead one to expect.
This is due in part to the ffrot that the weights include a
quantity of auxihary gear, such as air-bottles, etc., which form
a larger percentage of the whole weight in the case of smaller
engines. The fact of the weight per B.H.P. varying consider-
iably in different sizes renders this figure very unsuitable for
estimating purposes. The table below shews the weight per
cubic inch of bore' for the same range of engines, and although
here also the influence of the weight of the auxiliary gear may
THE PRINCIPLE OF SIMILITUDE
57
be noticed there is a much better approximation to constancy
in the figures.
inn- • . -
Ptdn
•—
'f^'
/r
SOO
2C
,;rj
c400
^
■'
^
^
)
5C
^
^\':?
^
d
^
^
^
-*-^
:^
4C
'^300
^
■^
,'
Q.
03
'\
,''
.-'
Ill
800 "2
vZOO
--;
-"
Q.
^
^^
%I00
OS'^
^
7rtfl
r,<'
7
o
V\
f
/
^
500
wa
b 10 lb 20 2b 30"
Bore oF Cylinderin Inches.
Fig. 20.
Table I
Weight in lb.-^[Bore'(inches)xNo. of Cylinders.]
1 Cylinder.
2 Cylinders.
3 Cylinders.
4 Cylinders.
Bore.
lb.
Bore.
lb.
Bore.
lb.
Bore.
lb.
9-0
15-0
9-0
12-0
12-6
7-1
16-5
6-2
10-6
13-2
10-6
8-5
13-8
7-1
18-1
61
11-4
12-0
11-4
8-3
15-0
7-1
22-4
6-1
12-6
11-2
12-6
7-8
18-1
6-7
—
—
13-8
11-2
13-8
7-8
20-5
6-5
—
—
15-0
11-2
15-0
7-8
—
—
— '
—
Table II below gives similar figures deduced from the
published weights of a line of four stroke land engines of the
58
DIESEL ENGINE DESIGN
crank-case type with trunk pistons, and all having three
cylinders. In this case the fly-wheel and auxiliary gear are not
included.
Table II
Bore in in 10-25 12 14 16
Stroke in in 15 18 21 24
Weight in lb. (bore 3x3) . 5-7 5-2 6-6 5-2
From these and other figures it appears that the crank-case
construction is not intrinsically much lighter than the "A"
frame type of construction, and that with the former any
economic advantages are traceable to the higher piston speeds
which become practicable when the crank-case is enclosed.
Similar figures are given in Table III for various types of
Marine Diesel Engines. The figures include the fly-wheel and
piping attached to the engine, but no auxiliary gear.
Table III
f.
Description.
Large two stroke. Cast iron columns similar
to steam engine practice ; water pumps
included
Large two stroke. Structure of trestle type
with through bolts ; water pumps included .
Large four stroke structure of trestle or crank-
case type with through bolts only, the H.P.
stage of compressor included . . . .
Large four stroke. Structure of stay-bolt type .
Trunk engines of crank-case type
Trunk engines of stay-bolt type
Weight in lb. ;-->
per^n. ' of bore', ^
X No. of cylind's.)
9-5 to 14
5-2
4-5
4-0
4-8
3-5
For preUminary estimating it is convenient to have an
approximate idea of the weights of the piston, connecting rod,
etc., of a proposed engine, and the following figures are a
rough guide to average practice. A designer wiU find it
convenient to make notes of similar figures for actual engines
with which he has had experience.
THE PRINCIPLE OF SIMILITUDE
59
Table IV
Name of Part.
Weight per in. '
of bore' lb.
Trunk piston (land engines) ....
0-18
Connecting rod „ „ ....
0-17
Crank-pin „ „ ....
0-036
One web „ ,, ....
0-036
Reciprocating weight (land engines) .
0-24
Revolving weight „ ...
0-18
Piston and piston rod (marine engine)
0-13
Crosshead „ ,,
0-13
Connecting rod „ „
0-24
Crank-pin „ ,,
0-046
One web „ ,, . .
0-046
Reciprocating weight ,, „
0-356
Revolving weight ,, „
0-236
The above figures are necessarily approximate only, as
they depend not only on the materials used and the views
as to design stresses, etc., held by individual designers, but
also on the bore to stroke ratio and the rules of insurance
societies.
Determination of Bore and Stroke. — Nowadays Diesel
Engines are generally built up of standardised units in com-
binations of 2, 3, 4, 6 and 8 cylinders, and by this means a large
range of sizes may be covered with a minimum of stock
patterns, jigs, etc. The choice of sizes of cylinders and the
numbers of different sizes stocked will of course depend on the
capacity of the factory and the estimated demand. Marine
engines are usually provided with 6 or 8 cylinders in the case of
four stroke engines and 4 or 6 cyhnders in the case of two stroke
engines, to facilitate starting and manoeuvring and also to keep
the size of cylinder and the aggregate weight within reasonable
limits. Having decided on the brake-power to be developed
by a certain standard cylinder the indicated power" must be
inferred, either from previous experience of engines similar to
the one proposed, by direct estimate of the various losses, or
with reference to published data of comparable engines. Typical
figures for mechanical efficiency have been given on page 27.
60 DIESEL ENGINE DESIGN
A very close approximation of the mechanical losses is obtain-
able as follows : —
1. Air compressor and scavenger losses may be reckoned to
be equal to the indicated power of these accessories
divided by a mechanical efficiency of 0-8.
2. Losses due to friction of piston rings about 5% of the
indicated power of the engine.
3. Lubricated friction loss is given very approximately by
the formula :
Horse-power lost in lubricated friction
_ 0-3 [D.LiVi^-s+n(A-KBL,)V,i-^]
550
Where
D =Diameter of crank-shaft in in.
Li=Aggregate length of journals and crank -pins in in.
Vi=Peripheral speed of journals in ft. per sec.
n =Number of cylinders.
A = Area of slipper guide in sq. in.
B =Bore of cylinder in in.
L2=Length of piston in in.
Vj^Piston speed in ft. per sec.
Assuming that the desired indicated power per cylinder is
now known, the former is related to the cylinder bore by the
formula : —
TXTT> V A 0-785.B2P.V B^P.V ...
I.H.P. per cylmder=^^^33^^^ = ^^^ (1)
for two cycle engines.
And half this amount for four cycle engines, both assumed
single acting. I
Where
P=Mean indicated pressure in lb. per sq. in.
V= Piston speed in ft. per minute.
The piston speed and mcOTi indicated pressure both vary
considerably in different cases, and existing practice in this
respect will be discussed later. Table V gives values of the
I.H.P. per sq. in. of bore^ for various values of the piston speed
and M.I.P. commonly adapted. The I.H.P. of a proposed
cylinder is found by multiplying the square of the bore by the
appropriate coefficient from Table V in the case of a two stroke
and J)y half that figure in the case of a four stroke cylinder.
THE PRINCIPLE OF SIMILITUDE
61
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62 DIESEL ENGINE DESIGN
Example : The I.H.P. of a 30 in. four stroke cylinder working
with a M.I.P. of 80 per sq. in. and piston speed of 900 ft. per
minute is:— 30»xO-857-^2=386 I.H.P.
Piston Speeds. — For trunk piston land engines of the open
type the piston speed usually varies from 600 ft. per minute
with very small engines of about 8 in. bore to 900 ft. per minute
with large engines of about 23 in. bore. A linear relation
between bore and piston speed between these limits is given
by the formula pig^.^^^ speed=460 + 17-6 (bore),
which represents good average practice.
With high speed forced lubricated trunk engines the practices
of the difEerent makers are not so consistent but appear on the
average to be based on an increase of about 20% above the
figures quoted above for slow speed engines. Some makers
adopt a uniform piston speed of about 800 ft. per minute for all
sizes, but greater uniformity of reliability would appear to be
obtained by a graduated scale of piston speeds, as indicated
above.
Mean Indicated Pressure. — ^Assuming that in every case the
engine is or may be required to run continuously at full load,
then the M.I.P. at rated full load is mainly dependent on the
cylinder bore in the case of four stroke engines for the following
reasons : —
1 . Effective cooling of the cyHnder walls, covers and pistons
becomes increasingly difficult as the size of the
cylinder is increased on account of the increased length
of stream lines through which the heat has to be con-
ducted.
2. As the cylinder bore is increased it becomes more difficult
to obtain an overload without smoke. With a 9 in.
cylinder it is possible to obtain an M.I.P. of 140 lb. per
sq. in. with an invisible exhaust and a rated M.I.P. of
105 lb. per sq. in. is peftnissible. With a 25 in. cylinder
120 lb. per sq. in. is about the limit, and it is found
advisable to restrict the rated M.I.P. to about 87 lb.
per sq. in. at nominal full load.
Table VI below gives the nominal full load M.I.P. for four
stroke engines for continuous running and provides for
occasional overloads of short duration, amounting to about
20%.
THE PRINCIPLE OP SIMILITUDE 63
Table VI
Bore of Cylinder, ina. 8 10 12 14 16 18 20 25 30
M.I.P. at rated
full load . .110 107 105 102 99 96 94 87 80
Table VI applies more particularly to land engines. For
large marine engines fitted with cooled pistons slightly higher
figures are sometimes used.
With two stroke engines another factor enters into the
question, viz., the efiiciency of the scavenging process and the
excess of air passed into the cylinder. With uncontrolled ports
and assuming a liberal surplus of air the maximum M.I.P.
obtainable is somewhere in the neighbourhood of 80% of that
obtainable with a four stroke engine of the same size. Apart
from the maximum pressm-e obtainable, a reduction of the
M.I.P. by 15%, as compared with a four stroke engine, places
both cycles on an approximate equality so far as the mean
temperature is concerned. A reduction of the figures given in
Table VI by 15% therefore appears rational and gives results
which agree fairly closely with published information of
conservative designs.
With supercharge devices, either in the form of valves in the
cover or valve controlled ports in the liner, the mean pressure
obtainable is only limited by the capacity of the scavenge pump
and the hmits imposed by the designer on the compression
pressure. With such engines M.I.P. of 170 lb. per sq. in. have
been obtained, and it is evident that with the supercharge
system high mean pressures are not necessarily associated with
high mean cycle temperatures, and no table can be laid down
for the M.I.P. permissible in such cases, but the following
formula is suggested : —
M.I.P. =Pi^-|- (2)
^ Po
where,
Pi= Permissible M.I.P. for a four stroke engine of the
same bore.
S=Stroke.
L=That amount of the stroke which remains to be per-
formed after the scavenge ports or valves are closed.
Po = Atmospheric pressure (absolute).
DIESEL ENGINE DESIGN
p=« Pressure (absolute) at the point of closing of the
scavenge ports or valves (usually 1 to 2 lb. in excess
of the pressure of the scavenge air).
Example :
Bore =20 in.
L=0-75S.
p = 19 lb. per sq. in. abs.
From Table VI, Pi=94 lb. per sq. in.
Therefore M.I.P. ^^^?'^f ^ ^^ =91 lb. per sq. in.
14-7 ^ ^
Both the above rules for two stroke engines assume that the
design of the scavenging apparatus is consistent with the
efficient expulsion of the exhaust gases.
Literature. — " Diesel Engine Cylinder Dimensions." —
Engineering, September 26th, 1913.
Richardson, J., " The Development of High Power Marine
Diesel Engines." — Junior I. E., April 20th, 1914. This paper
contains a great deal of information regarding the dimensions,
weights, and capacities of the leading types of Diesel Engines.
CHAPTER V
CKANK-SHAPTS
Material. — For Marine Diesel Engine crank-shafts the usual
material is open hearth steel having a tenacity of 28 to 32 tons
per sq. in. and minimum elongation of 25 to 29% on 2 in.,
the lower minimum elongation being associated with the
higher tenacity.
For stationary engines it is more usual to employ steel of a
tenacity of 34 tons and upwards, specif3dng a minimum
elongation of 25% in 2 in.
Centres of Cranks . 2-65 Thickness of Webs .
Diameter of Shaft . 0-63 Width of Webs
' Length of Crank-pin 0-63
Fig. 21.
General Construction. — The majority of Diesel Engine
crank-shafts are turned from solid forgings. When the ratio
of stroke to bore exceeds about 1-8 a built-up shaft is some-
times used, and typical proportions are given on Fig. 21,
the unit being the bore of the cylinder.
Fig. 22 shews a solid forged crank-shaft for a two cylinder
stationary engine.
For four stroke stationary engines the shaft is usually of
one piece when the number of cylinders does not exceed four.
F 65
66
DIESEL ENGINE DESIGN
For six cylinder stationary engines the usual practice is to
divide the shaft into two sections, arranging the cam-shaft
drive in the centre. Occasionally the one piece arrangement
is adopted with the cam-shaft drive at one end, preferably the
fly-wheel end. This makes the neater arrangement, and as no
difficulties appear to occur in practice, probably the only dis-
advantage is the expense of replacing the whole shaft in the
event of failure. The turning of long crank-shafts offers no
difficulties provided a modern crank-shaft lathe is available.
For four stroke marine engines of six and eight cylinders the
usual arrangement is two strictly interchangeable sections of
shaft. With two stroke marine engines the cylinders are
arranged in pairs, with a section of shaft to each pair, the
cranks of which are placed at 180°. This arrangement complies
IQ
y
•y
-.^
Fio. 22.
with the requirements of good balance and equal division of
impulses, and the fact of both cranks of a section being in a
plane facilitates both forging and machining. To secure inter-
changeabihty the number of bolts in each coupling should be
either equal to or a multiple of the number of cylinders.
Arrangement of Cranks and Order of Firing. —
1. Two cylinder four stroke engines.
The cranks are commonly arranged on the same centre and
the cyUnders fire alternately A equal intervals, thus sacrificing
balance to equal spacing of impulses. Arranging the cranks at
180° does not very materially affect the degree of uniformity
and has the advantage of balancing the primary forces, but the
primary couples are unbalanced. The latter would appear to
be the lesser evil.
2. Three cylinder four stroke engines.
Cranks at 120° and firing periods follow at equal intervals of
CRANK-SHAFTS 67
240°. Primary and secondary forces are balanced, but un-
balanced primary and secondary couples exist. The two latter
do not appear to be very serious so far as their effects in
producing vibrations are concerned.
3. Four cylinder four stroke engines.
The most common arrangement is to have all the cranks in
one plane, the inner pair of cranks being on the same centre
and the two outer cranks at 180° to them. Ignitions follow at
intervals of 180°. The primary forces, and both the primary
and secondary couples, are balanced, but the secondary forces
are completely unbalanced. Vibration troubles are common
with this arrangement. By arranging the cranks as for a two
cycle engine {q.v.) both primary and secondary forces can be
balanced at the expense of unequal spacing of ignitions.
4. Six cylinder four stroke engines.
In principle each half of the shaft represents the optical
image of the other half as seen in a mirror placed at the centre
of the engine in a plane at right angles to the centre line of the
shaft, the cranks in each half being at 120° to each other, as
for a three cylinder engine. The primary and secondary
couples generated by each half of the engine mutually cancel
each other, so that complete balance is obtained so far as
primary and secondary forces and couples are concerned.
5. Eight cyhnder four stroke engines.
The same principle of equal but opposite handed shaft halves
apphes to this case also. Each half consists of four cranks, the
outside pairs of which are at 180° and the planes containing
these pairs being at right angles. (See Fig. 23, which also
shews an alternative arrangement.)
Each half of the engine is balanced for forces and the two
halves balance each other for couples. Similar arrangements
are possible for higher even numbers of cylinders.
6. Four cylinder two stroke engines.
The arrangement for cranks is the same as that described for
one half of the shaft for an eight cylinder four stroke engine.
7. Six cylinder two stroke engines.
The arrangement consists of three pairs of cranks, the
individuals of each pair being at 180° and the planes contain-
ing the pairs being at 120° to each other. Secondary couples
only are out of balance.
68
DIESEL ENGINE DESIGN
8. Eight cylinder two stroke engines.
Similar to six cylinder engine, but planes containing pairs of
cranks at angles of 45°. Primary and secondary couples out
of balance. With two stroke engines, owing to the fact that
no two cranks are on the same centre, the order of firing is
determined by the angular position of the cranks only. With
four stroke engines, on the other hand, having four, six or eight
cylinders for every firing point there is a choice of two cylinders
and the orders of firing generally adopted are based on the
principle of placing consecutively firing cylinders as remote as
4 Stroke Engines.
N?of
Cylinders
Arrangement of Cranks
Order of firing
-JTUTL.
03 ^h^cDl
12 12
3^.
132
10):
1243
3-4icyz-5
153524
16284735
1526 8473
2 Stroke Engine s.
,^.
123
'^'
1423
145236
44
57
16472538
Figs. 23, 24.
CRANK-SHAFTS 69
possible, so as to avoid local accumulation of elastic strain due
to the reaction at the bearings. The usual arrangements and
sequences of cranks, and also the order of firing, are shewn
diagrammatically on Figs. 23 and 24 for four stroke and two
stroke engines respectively.
Lubrication. — ^The question of the lubrication of Diesel
Engines is partly influenced by the adoption of crossheads
and guides distinct from the piston. So long as the trunk
piston is used forced lubrication is at a disadvantage, on
account of the difficulty which is experienced in pre-
venting the splashing of oU on to the cylinder walls and
disadvantages arising from the mixing of carbonised oil from
the cylinder with that used for the bearings. The use of a
crosshead, on the other hand, enables the cylinder to be isolated
from the crank chamber, and forced lubrication can then be
used with the same success which attends its application to high
speed steam engiaes. To obtain the maximum benefit from
this system it is necessary for the crank-case to be carefully
designed to prevent loss of oil, either in the form of splash or
impalpable mist, and also to prevent the ingress of grit. It is
a mistake to suppose that a copious supply of lubricant under
pressure necessarily eliminates wear. Only by the absolute
exclusion of grit from the crank-case and from the entire
lubricating system can wear be reduced to a minimum. For
this reason it would seem desirable for all lubricating oil pumps,
filters, etc., to be located within the crank-case itself. An
exception to this rule may advantageously be made in the case
of oil coolers if these are fitted. The presence of cold bodies
within the crank-case results in the condensation of water,
which mixes with the oil unless special arrangements are made
to cope with this difficulty.
So far as economy of lubricating oil is concerned, ordinary
ring lubrication for the main bearings and the centrifugal
banjo arrangement for the big ends leave nothing to be desired.
With suitable arrangements for filtering the oil which is drained
from tiie crank-pit, and using over and over again, the nett
lubricating oil consumption is readily kept below 0-002 lb.
per B.H.P. hour (trunk piston engines). When forced lubri-
cating is appUed to high speed trunk engines the consumption
is frequently higher. Efficient use of non-forced lubrication
necessitates certain special features in connection with the
crank-ehaft. Rings are turned on the latter at each end of
70
DIESEL ENGINE DESIGN
each journal to throw the squeezed out oil into suitable catcher
grooves in the bearing brasses, whereby it is returned to the
oil well instead of being thrown o£E by the crank webs. As
these oil throwers have been shewn in some cases to weaken
the shaft at its already weakest point, they should be designed
BO as not to interfere with a good radius between the journal
and the web. Fig. 25 shews a section through such an oil
thrower.
Fig. 26 shews a crank fitted with centrifugal banjo lubricator.
The oil hole leading to the surface of the crank-pin is preferably
Fig. 25.
Fig. 26.
drilled at an angle of about 30° in advance of the dead centre,
so that the upper connecting rod brass receives a supply of oil
just before the ignition stroke.
With forced lubrication oil throwers and catchers are not
usually fitted, but the shaft requires to be drilled to conduct
oil from the journals to the crank-pins. Two systems of drilling
are shewn in Figs. 27 and 28.
Details. — (1) Webs. — Various types of solid forged webs are
shewn in Figs. 29, 30 and 31. The two ends of the straight-
sided webs are sometimes turned from the journal and crank-
pin centres respectively. Weight can be reduced slightly by
turning the two ends at one setting from a centre midway
between these two points. Triangular-shaped segments are
CRANK-SHAFTS
71
usually turned off the projecting corners of the webs, and this
reduced weight facilitates feeling the big ends of the connecting
rods and gives more clearance for the indicating gear. Balance
Fig. 27.
Fig. 28.
weights are frequently fitted to one and two cylinder engines
to balance the revolving weight of the crank-pins, the big ends,
and the otherwise unbalanced portion of the crank webs. The
chief difficulty in designing a balance weight is usually to get a
Fig. 29.
Fig. 31.
sufficiently heavy weight in the space available. The magni-
tude of the balance weight required is equal to the weight to be
balanced (i.e. weight of crank-pin plus about 0-65 of the total
weight of the connecting rod and about half the weight of the
72
DIESEL ENGINE DESIGN
webs) multiplied by the radius of the crank and divided by the
radius measured from the centre of the shaft to the centre of
gravity of the balance weight. The problem thus resolves
itself into a matter of trial and error. Various
modes of securing balance weights are illus-
trated in Pigs. 32, 33 and 34. The bolts, or
other form of attachment, should be suffi-
ciently strong to carry the centrifugal force of
the weight with a low stress.
(2) Couplings. — The couplings connecting
the sections of a shaft are made with spigot
and faucet joints, the spigots being turned
off after the bolts have been fitted. The
bolts belonging to the coupUng to which a
gear-wheel is fitted are usually made of
additional length, and used to secure the
wheel. If separate means of securing the
wheel are used the interchangeability of the
sections of shaft is prejudiced. This is of small importance
where land engines are concerned. With marine engines,
where it is desirable to carry a spare section of shaft, the
Fig. 32.
Fig. 33,
Fig. 34.
latter should be provided with any keyways, extra bolt-holes,
etc., requisite to enable it to replace any section of the shaft
in the event of failure, with a minimum of fitting.
When heavy fly-wheels are fitted, as for instance with dynamo
CRANK-SHAFTS
73
drives, an outer bearing is sometimes placed between the fly-
wheel and the driven shaft. The coupling used to connect the
projecting end of the crank-shaft to the drive may conveniently
be of the common cast iron flanged type, provided with a
shrouding to cover the nuts and bolt-heads. See Fig. 35.
Fig. 35.
Occasionally this outer coupling is forged with the shaft,
and if precautions are taken to ensure that the coupling is free
of all bending moment it may be proportioned to the twisting
moment only. This arrangement is a convenient one when the
cam-shaft drive is at the fly-wheel end of the engine, as the
small diameter of the coupling enables the crank-shaft gear-
wheel to be passed over the coupling in one piece. See Fig. 36.
Fig. 36.
(3) Air Compressor Cranks. — ^When the air compressor is
supplied by another manufacturer the compressor crank is
usually designed by the latter. This arrangement is not
always quite satisfactory as the apparatus by means of which
the compressor is driven at the shop test may be more favour-
able for satisfactory running than the arrangements which are
74
DIESEL ENGINE DESIGN
made for the reception of the compressor on the engine. In
any case close co-operation should exist between the two
manufacturers to produce a satisfactory job between them.
The main points to be insisted on are rigidity and truth of
alignment. Figs. 37 and 38 illustrate two satisfactory designs
of crank.
(4) Scavenger Cranks. — ^Whenthe air compressor is driven
by an overhung crank extending from the scavenger crank-
shaft, the latter must be made far stiffer than would be required
from considerations of strength alone, in order to keep the
deflection of the overhung crank within small limits. In this
case it is not unusual to make the scavenger crank the same
diameter as the main crank-shaft. When two scavengers are
- ^=)
v_
Fig. 37.
m
Fig. 38.
provided the cranks are placed at 90° to equahse the discharge
of air.
Proportions. — In view of the fact that the straining actions
causing fa;ilure of crank-shafts are mainly due to bending
moments caused by unequal level of the bearings, the most
consistent results are obtained when discussing the proportions
found in practice by expressing aU dimensions in terms of the
cyhnder bore. As regards marine crank-shafts, the designer
has little latitude, as minimum»dimensions are fixed by the
rules of insurance societies. To evaluate these rules the
distance between centres of cylinders must be determined,
and this figure is usually about twice the cylinder bore in the
case of four stroke engines and about 2 to 2-4 times the bore
in the case of two stroke engines. The diameter of the shaft
usually works out at about 0-62 for four stroke and 0-65 for
two stroke engines. These approximate figures are merely
quoted here for comparison with those relating to land engines.
CRANK-SHAFTS
75
Proportions of Four Stroke Land Engine Crank -shafts. —
Typical proportions for a slow speed engine are given below,
the unit being the cylinder bore :—
Diameter of journal and crank-pin . . . . 0-53
Length of journal 1-15
Length of crank-pin 0-53
Width of web 0-80
Thickness of web 0-28
Proportions of an exceptionally strong shaft are given
below : —
Diameter of crank-pin and journal
Length of journal
Length of crank-pin
Width of web ....
Thickness of web ....
0-57
0-88
0-70
0-92
0-32
Mr. P. H. Smith, in a paper read by him before the Diesel
Engine Users' Association, July, 1916, recommends the follow-
ing proportions applying to 34-ton steel on the understanding
that certain precautions are taken to keep the bearings in line.
Diameter of crank-pin and journal . . 0-525 to 0-54
Length of journal 0-75 „ 0-80
Length of crank-pin 0-524 ,, 0-54
Thickness of web 0-32 minimum
Mr. Smith points out that Diesel Engine crank-shafts almost
invariably fail at the webs, and the thickness he proposes for the
latter is the greatest the author has found in practice.
According to Mr. Smith's minimum figure for the shaft
diameter and web thickness, and taking the width of the web
to be 0-8 of the cylinder bore, the relative strengths of the
journal and the web to resist bending are as
7rx0-5253 0-8x0-322
^__to — g
= 1 : 0-96,
so that a shaft based on these proportions will be of nearly
equal strength throughout if the effects of radii and changes
of shape are neglected. Crank-shafts for two stroke land
engines are generally of slightly larger diameter than those
for four stroke engines. Different examples give figures vary-
ing from 0-55 to 0-59 of the cylinder bore. Great difference of
opinion exists as to the size of the radii between journals and
76 DIESEL ENGINE DESIGN
crank-pins and, crank webs. A radius of 0-07 of the shaft
diameter is good average practice, though some designers
prefer 0-15 and others are satisfied with 0-04.
Calculation of Stresses in the Crank -shaft. — It is as well to
state at the outset that the problem of determining the stresses
in a multi-crank-shaft is rather laborious, if done conscientiously
and actual designs are more often than not based on experience
pure and simple, without reference to comparative calcula-
tions other than of simple proportion. Supposing a suitable
analysis to have been made for a correctly aligned shaft, the
whole calculation would require to be revised before the results
could be appUed with accuracy to the case of a shaft of which
the bearings were at different heights owing to the unequal
wear of the white metal or flexure of the bedplate. This latter
consideration is of itself valuable in emphasising the need of
haassive foundations where land engines are in question and
the desirabihty of providing an extremely rigid framework in
marine designs. Probably the severest condition with which
a Marine Diesel Engine has to contend is the deflection of the
hull due to variations of cargo loading, etc. A stiff seating in
way of the engine is doubtless helpful, but the surest way of
avoiding deflection troubles is to design the engine framework
in the form of a deep box girder. Apart from other considera-
tions, this enables the engine to be erected in the shops once
and for all, without the necessity of re-bedding the shaft after
installation in the ship. The component parts of the crank-
shaft, viz., journals, crank-pins and webs, are subject to
bending and twisting actions which vary periodically as the
shaft revolves. In the past it has been customary to compute
equivalent bending or twisting moments corresponding to the
calculated co-existing bending and twisting moments and to
proportion the shaft accordingly. Recent experiments by
Guest and others indicate that steel under the influence of
combined bending and twistii^ begins to fail when the shear-
ing stresses, as calculated from the formula quoted below,
attain a definite value (about 12,000 lb. per sq. in. for very mild
steel under alternating stress) which is independent of the
relative amounts of bending and twisting.
Maximum shear stress = Jv'4/B^+fn^
Where /o=Normal stress due to bending.
/, = Shear stress due to twisting.
CRANK-SHAFTS 77
The equivalent twisting moment which would, give the same
shear stress as the maximum shear stress due to the combined
action of the actual bending and twisting moments is given by
Where Tj;= Equivalent twisting moment.
T= Actual twisting moment.
B =Bending moment.
A good approximation to the twisting moment at any point
of the shaft at any degree of revolution is obtained by combin-
ing in correct sequence the twisting moment curves correspond-
ing to all cylinders " for'd " of the section under consideration.
In future the terms " forward " and " af t " wUl be used to
denote the compressor and fly-wheel end of the engine respec-
tively regardless of whether the engine under consideration is
of marine or land type. The negative twisting moment due to
mechanical friction of the moving parts is almost always
neglected. That due to the compressor is sometimes allowed
for. When dealing with the stress in a crank-pin it should be
borne in mind that the twisting moment due to any cylinder
is not transmitted through its own crank-pin. For example,
if the cranks are numbered as usual from the compressor end,
the twisting moment in No. 3 crank-pin is that due to cylinders
Nos. 1 and 2. The calculation of bending moments is by no
means straightforward, and the methods adopted form the
distinguishing features of the systems of crank-shaft calculation
described below.
Fixed Journal Method of Crank-shaft Calculation. — ^The
assumption underlying this method is that each journal is
rigidly fixed at its centre and that the section of shaft between
two jomrnals may therefore be treated as a beam encastr6 at
its ends. The assumption of fixed journals would be true for
a row of cylinders all firing at the same time. For ordinary
conditions, however, the assumption would only hold good if
the bearings had no running clearance and were capable of
exerting a bending effect on the shaft by virtue of their rigidity.
Apart from the fact that the construction of bearings and
bearing caps does not suit them for this heavy duty, examina-
tion of the bearing surface of well-worn bearings reveals no
trace of such cornering action and justifies the view that the
bearings merely fulfil their proper functions of carrying thrust
in one direction at a time.
78 DIESEL ENGINE DESIGN
Free Journal Method. — ^With this system each crank is
supposed to be loaded at the centre of the crank-pin and sup-
ported freely at the centre of the journals, so that the maximum
bending moment occurs at the centre of the crank-pin. This
assumption would be approximately true for a single cylinder
engine if the weight of the fly-wheel and the influence of out-
board bearings are neglected. With this method the twisting
moment is of very secondary importance, in many cases almost
negligible. Comparing this system with that described above,
it will be seen that both involve the construction of twisting
moment diagrams, though the accuracy of the result is of less
importance in the case of the free journal method.
In the following articles a four throw crank-shaft will be
investigated on somewhat different lines, with a view to
eliminating as many unjustifiable assumptions as possible.
Stress Calctjlation for a Four Throw Crank-shaft
Data :
Type of engine four stroke
Number of cylinders four
Bore of cylinders 10 in.
Stroke 15 in.
Revolutions per minute 300
Connecting rod 5 cranks long
Maximum pressure at firing dead centre, 500 lb. per sq. in.
Diameter of journals and crank-pins . . 5-25 in.
Length of journals 8 in.
Length of crank-pins 5:5 in.
Thickness of webs 3-25 in.
Width of webs 9 in.
Centres of cyUnders .... 8+5-5+6-5=20 in.
Weight of piston 170 1b.
Weight of connecting rod . . • . . . . 190 lb.
Weight of crank-pin . .% 33 lb.
Weight of unbalanced parts of two crank-webs . 110 lb.
The method employed in the following investigation is to
calculate the values of the forces acting on the shaft when one
crank is at its firing top dead centre. The reactions on the
bearings will be calculated on the assumption that the centres
of the journals remain level and all loads will be treated as
concentrated. The effect of unequal level of bearings will also
CRANK-SHAFTS 79
be investigated. In computing the forces acting on the crank-
shaft the dead weight of the latter, and also that of the running
gear, fly-wheel, etc., will be neglected and the effect of the air
compressor will not be considered, nor will the small exhaust
pressure remaining in the cylinder which completes its exhaust
stroke at the same instant that the cylinder under considera-
tion begins its firing stroke, so that the forces to be dealt with
are : —
(1) Those due to cylinder pressure.
(2) Centrifugal force of revolving parts.
(3) Inertia force of reciprocating parts.
These will now be calculated : —
Weight of revolving part of connecting rod,
0-65x190 124 1b.
Weight of unbalanced part of crank webs . . 110 lb.
Weight of crank-pin 33 lb.
Total weight of unbalanced revolving parts 267 lb.
Weight of reciprocating part of connecting rod,
0-35x190 66 1b.
Weight of piston 170 lb.
Total weight of reciprocating parts . . 236 lb.
Centrifugal acceleration in the crank circle is : —
, /27r.300V 7-5 „,^,^
w2r==( — ^fT—J X Y^ =615 ft. per sec.^
Therefore centrifugal effect of revolving parts
?^l><^ = 51001b.
g
Inertia effect of reciprocating parts at top dead centre : —
236x615 ,, K.nAiu
X l+=5400 lb.
g
Inertia effect of reciprocating parts at bottom dead centre : —
?5^^^X(l-i) = 3600lb.
g
Combined centrifugal and inertia effect at top dead centre
=5100-1-5400 = 10,500 lb. upwards.
Combined centrifugal and inertia effect at bottom dead
centre
= 51004-3600 = 8700 lb. downwards.
80
DIESEL ENGINE DESIGN
Maximum load due to cylinder pressure
=0-785 X 102 X 500 = 39,000 lb.
Resultajit downward effect of pressure, centrifugal force and
inertia force at firing dead centre
= 39,000-10,500 = 28,500 lb.
Method of Calculating the Reactions at the Bearings. —
Let A 1 (Fig. 39) represent the centre of the crank-shaft,
A C E G I being the centres of the journals and B D F H
those of the crank-pins.
Rg R4 Re Rs a-re the applied forces due to cylinders 1, 2, 3, 4
respectively.
RjRgRjR^Rg are the reactions at the bearings unknown
in magnitude and direction. Under the influence of these
forces the centre line of the shaft assumes some deflected shape,
Cyin. CyirZ. Cyl':3. CylTA.
/?? f{4 /?5 Rg
I
Fig. 39.
Rg
and the deflection at any point above or below the straight
line joining A 1 is equal to the sum of the deflections at the
same point which would be produced by each of the forces
R2R3R4R5R6R7R8 acting alone, supposing the shaft
supported freely at A and 1 . At C E and G the sum of these
deflections must be zero if the bearings are level (ignoring the
effect of running clearance). Let Cg Cg and g^ be the deflections
at C E and G due to unit load applied at B (the position of Rj),
assuming the shaft support||d freely at A and 1.
Cgesandga'
0464 and g4
Cgegandgs
etc., etc.
Caegandgg
are the corresponding deflections at C E and G
due to unit load applied at C D E, etc.
The values of Cj Ca and gg, etc., are readily found by the usual
formulae for the deflection of beams.
CRANK-SHAFTS 81
Now since the total deflection at C E and G is zero, the
following equations hold good : —
R2 C2+R3C3 + R4C4 + R5C5+ReC6 + R,C,+R8C8 = (1)
R2e2+R3e3+R4e4+R6e5+R6e6+R,e,+R8e8=0 (2)
R2 g2+R3 g3+R4 g4 + R6 gs+Re g6+R7 gT+Rg g8=0 (3)
In these three equations the only unknown quantities are
R3 Rj R7, which can therefore be determined. The remaining
unknown reactions R^ and Rg are found by equating moments
about 1 and A respectively. In solving the above equations
downward forces and deflections will be considered positive
and upward forces and deflections negative.
Determination of C2e2g2, etc. — In determining these
constants it will be assumed that the shaft deflects under load
as though it were a cylindrical beam of the same diameter as
the crank-pins and journals. Considering that the webs of a
Diesel Engine crank are short, the presumption is probably
not inaccurate, but it would be interesting to see this point
investigated, as it could readily be, by means of models. So
long as bearings at constant level are assumed, the actual
value of the deflection is of no importance, as the method of
calculation depends only on the relative deflection at the
different points considered. The problem therefore resolves
itself into finding the deflected form of a beam freely supported
at each end under the influence of a concentrated load placed
anywhere between the supports. This may be done by treat-
ing each end of the beam as a cantilever. The deflection of the
end of a cantilever carrying a load at the end is given by : —
W 1'
Deflection at the end of cantilever in in. = „'
— 3 EI
Where W=Loadinlb.
E = 30,000,000 lb. per sq. in. (for steel).
I=Moment of inertia (transverse) of the section of
beam in.
For the shaft under consideration : —
I=Zx5-25* = 37-3in.*
64
Fig. 40 shews the values of the deflection at various fractional
points in the length of the cantilever, the deflection at the end
being unity. These results are applied to the case of a beam
as follows : —
Let AB be a beam .(Fig. 41) supported at A and B and
82
DIESEL ENGINE DESIGN
carrying a load W^ at any point C. The reaction at A is equal
OB
to Wj j^. Let this be denoted by Wj. At A erect a perpendi-
cular AD equal to some convenient scale to the deflection of the
cantilever AC, due to the load W 2 at its end. Draw the deflected
shape of this cantilever (DC) by means of the proportions given
re 'A fa k % % ^e
-^
-^
r»
<o"
!i
00
s?^
s>
•s^
«5
c\
1
2
1
10
CM
Fig. 40.
in Fig. 40. Proceed similarly with cantilever CB, obtaining
the deflected shape C F. Join D and F, then the deflection of
the beam at any point X is the vertical intercept " x " shewn
in the figure.
Application to the Case in Hand. — The constants c^e^gi
being the deflections at various points on a beam due to unit
load applied at other various points, are independent of the
system of loads and wiU therefore be dealt with before special
A
1
X
D
<
c
X ""--.^
F
W3
Fig. 41.
cases of loading are consiJfered. In order to obtain manage-
able figures, the deflection will be reckoned in thousandths of
an inch and the unit load will be taken as 10,000 lb. Deflection
of cantilevered portion of shaft in thousandths of an inch is
given by:- W.(A)« _ W.(3^)3
3x30x37-3
Where W=Loadinlb.
Z=Length in in.
3360
CRANK-SHAFTS
83
The diagram Fig. 42 shews the process of determining
^2 ©2 §2! etc., in particular : —
Unit load at B = 10,000 lb., AB = 10", BI = 70"
-D ,. ^ . 10,000x70 „„^^,,
Reaction at A = — '—— =8750 lb.
Hi)
T 10,000x10 ,„.„,,
„ 1= g„ =1250 lb.
Deflection of cantilever AB= — -— - — =2-6^-^°^
odOU
^-^
^
f
\
^
^
y
///
/
N
■v ITj
^
-^Z
^
/]
/
^'^
^
4
i
/ ^
V
/
ABCOeFGHI
Fig. 42.
Deflected shapes of cantilevers drawn by plotting ordinates
from the proportion given in Fig. 40.
By scaling the diagram : —
C2 = 30-5 e2=35 g2=20-5
Since the deflection at G, due to a load at H, is the same as
that at C, due to the same load at B, and so on, therefore : —
g8 = 30-5 68 = 36 C8=20-5
By the same methods (see Fig. 42) : —
C3=51 e3=62 g3=39
g, = 51 e, = 62 c, = 39
c^ = 65 64=88 g4=55
g6 = 65 66 = 88 C6 = 55
C6 = 65 65 = 95 g6 = 65
84
DIESEL ENGINE DESIGN
These values could be computed with rather less trouble and
with greater accuracy by the formulae for the deflection at any
point in a beam due to a load at any other point (see Morley's
" Strength of Materials "). In more complicated cases, with a
larger number of unknown reactions, the constants c, e, g
require to be ascertained very accurately, otherwise large
errors arise.
Z-85 0-87 087 -105
Fio. 43.
PJ
Rgi
The conditions of loading will now be considered.
Case I. Crank 1 on Firing Dead Centre. — ^The magnitudes
and directions of the applied forces are shewn in Fig. 43.
Hence : —
R2C2 = 2-85x30-5 = 87-0 R
Ra 62 = 2-85x35 =99-5 R
R2g2 = 2-85x20-5 = 58-5 R4g4=0-87 x 55=48-0
04=0-87x65=56-5
4 64=0-87x88 = 76-5
Rj 06=0-87x55 = 48-0 Rg C8=-l-05 x20-5= -21-5
Rg 66=0-87x88 = 76-5 Rg e8=-l-05 x35 =-37-0
Rgg5=0-87x65 = 56-5 R8g8=-l-05x30-5=-32-0
From which
Ra C2+R4 C4+R6 C6+R8 08=170-0
R2 e2+R4 e4+R6 ee+Rs 68=215-5
R2g2+R4g4 + R6g6+R8g8 = 131-0
Substituting these values in equations (1), (2), and (3), we
obtain : — *
51 R3 + 65 R6 + 39 R7= -170-0 (4)
62 R3+95 R5 + 62 R,= -215-5-
39 R3+65R5 + 5IR, = -131-0-
-(5)
-(6)
From which
R3= -2-295 R5= -1-402 R,= -0-965
Equating moments about A : —
8 Rs + 7 Re-F6 R7 + 5 R6+4 R6 + 3 R4-K2 R3-hR2=0
CRANK-SHAFTS
85
.-. 8 R9 = 7x 1-05-6 x0-965-5x0-87+4x 1-402-3 xO-87
+ 2x2-295-2-850
Whence Rg=0-118.
Equating moments about I : —
8Ri + 7R2 + 6R3+5R4+4R5 + 3R6 + 2R,+R8=0
.-. 8 Ri= -7 X 2-85 + 6 X 2-29-5x0-87 + 4x1-402-3x0-87
-2x0-965 + 1-05
Whence Ri= -1-079.
Knowing all the forces, the bending moment at A B C, etc.,
can now be tabulated thus : —
Point.
Moments.
B.M. in
in. lb.
A
B
C
D
E
F
G
H
I
10,510x10
10,510x20-28,500x10 ....
10,510x30-28,500x20+22,950x10 .
10,510 X 40-28,500 X 30+22,950 X 20-8700
XlO
-2440x30 + 10,500x20-9,650x10 .
-2440x20 + 10,500x10.
-2440X10
105,100
-74,800
-25,200
-62,600
30,300
56,200
-24,400
Since the bending modulus of the shaft is
32
x5-253 = 14-2in.3
Therefore,
Maximum stress due to bending (occurring in No. 1 crank-
pin at top firing centre) is equal to : —
1^0^7,400 lb. per in.^
Bending stress, according to fixed journal method : —
W.L 28,500x20 __ ,,
rz^ 8x14-2 =50201b.sq.m.
According to free journal method : —
„, 28,500x20 ,„„^„,,
Stress = A^iAo =10.040 lb. sq. m.
86
DIESEL ENGINE DESIGN
Case II. No. 2 Crank on Firing Dead Centre.— The
magnitudes and directions of the applied forces are shewn in
Fig. 44.
R2=0-870 R4=2-850 R6=-l-05 R8=0-870
R2C2=0-87x30-5 = 26-5 R^ C4=2-85x65-0 = 185-0
Ra 62=0-87 x36-0 = 30-5 R^ 04=2-86 x88-0 = 251-0
R2g2=0-87x20-5 = 18-0 R4g4=2-85x55-0 = 156-5
Rg 06= -1-05 X 55-0= -58-0
Rg 66= -1-05x88-0= -92-5
R5gg= -1.05x65-0 = -68-0
Rg 08=0-87x20-5 = 16-0
Rg 68=0-87x35-0 = 30-5
R3gg=0-87x30-5 = 26-5
From which
Rg C2+R4 C4+R6 C6+R8 C8 = 171-5
Rg Og+Ri e4+R6 eg+Rg 68 = 219-5
0-87
Z-85
-l-OB
0S7
R, A?3 ffg /?7 Rg
Fig. 44.
The three equations for determining R3R5R, are there-
fore : —
51 R3+65 R6i39 R7= -171-5
I
-219-5
62Rj-|-95Rs-f62R,
39R3-f65RB-|-5lR7=-133
from which
R, = -2-005 Rs=-l-788 R, = 1-204
Taking moments about A and I,
Ri=-0-161 Rj =
-0-790
CRANK-SHAFTS
87
from which the following figures for the bending moments are
obtained : —
Point.
Moments.
B.M. in
in. lb.
A
B
1610x10
16,100
C
1610x20-
-8700x10.
,
-54,800
D
1610x30-
-8700x20-1-20,050x10
.
74,800
E
1610x40-
10 .
-8700x30-1-20,050x20-
-28,500 X
-80,600
F
7900X30-
-8700x20-12,040x10
,
-57,400
G
7900x20-
-8700x10.
71,000
H
7900 X 10
.
79,000
I
•
■
Maximum bending stress ,' „ =5670 lb. sq. in.
° 14-2
Considerably less than in Case I.
Case III. Loading as in Case I, but bearings C and G
supposed worn down 20 thousandths and bearing E 25
thousandths of an inch below the level of the line joining A I.
The difference in height of the bearings is very simply
allowed for by putting 20, 25, and 20 respectively on the right-
hand side of equations (1), (2), and (3) instead of zero.
The equations for determining RgRjand Rjthen become :—
51 R3-t-65 R5-h39 R7 = 20-
62 R34-95 Rs-l-62 R7 = 25-
39 R3-I-65 R5-f51 R,=22-
-170-0=-150-0
-215-5=-190-5
-131 =-109
The following values are obtained for the reactions at the
bearings : —
Ri=-1-412 R3=-l-610 R5=-2-135 R7=-l-820
R, = -0-202
88
DIESEL ENGINE DESIGN
The
bending moments are as follows : —
Point.
Moments.
B.M. in
in. lb.
A
B
c
D
E
F
G
H
I
14,120x10
14,120x20-28,500x10
14,120x30-28,500x20 + 16,100x10 .
14,120x40-28,500x30 + 16,100x20-8700
XlO
2020x30 + 10,500x20-18,200x10
2020x20 + 10,500x10
2020x10
141,200
-2,600
146,000
-55,200
88,600
145,400
20,200
Maximum bending stress
146,000
14-2
= 10,250 lb. sq. in.
Comparing the above figures with those obtained in Case I,
it will be seen that the maximum stresses have been increased
to the extent of about 35% by the difference of level of the
bearings. In view of the uncertainty which exists with regard
to the deflection of cranked shafts the above figures are not
strictly reliable, but the writer is of opinion that they under-
rather than over-estimate the stresses.
Conclusions. — 1. The value of the bending moments at a
crank-pin on the top firing centre is greater for a crank-pin
situate at one end of the shaft than that at one nearer the
centre of the shaft.
2. The bending moments at certain crank-pins and journals
may be as great or greater than the bending moment at the
crank-pin which is receiving the greatest applied load.
3. No general rule for the be)|ding moment at a Diesel Engine
crank-pin or journal can represent the true state of affairs,
but every different arrangement of cranks and number of
cylinders requires to be investigated individually.
4. Difference of level of the bearings, due to wear or other-
wise, gives rise to greatly increased bending moments, which
can be calculated approximately in the manner described.
The methods of calculation which have been illustrated in
CRANK-SHAFTS 89
this chapter can be applied to cases involving any number of
cranks, and the effects of fly-wheels, the rotors of electric
generators, outboard bearings, etc., can be included. When
the number of cylinders is three or less the weight of the fly-
wheel is considerable, and cannot therefore be ignored. In
these cases allowance should be also made for the practice of
packing the outward bearing above the level of the engine
main bearings. For engines of four cylinders and over, the
weight of the fly-wheel and the presence of outboard bearings
can probably be neglected with safety. An extended treat-
ment of this subject will have to be reserved for a future
occasion.
The labour involved in solving simultaneous linear equations
increases as the square of the number of unknowns. For a
description of a machine devised to do this work mechanically,
see the " Treatise on Natural Philosophy," vol. i., Kelvin and
Tait.
Graphical Determination of the Twisting Moments. — ^In
the processes described below the following approximations
have been made : —
1 . The negative twisting moments due to the air compressor
at the forward end of the engine have been neglected. These
moments are small in comparison with the moments due to the
working cylinders, and being opposite in direction to the
maximum moments tend to reduce the latter by a small
amount.
2. Moments due to the dead weight of the revolving and
reciprocating parts have also been neglected. In a very large
engine it would be advisable to take these into consideration,
as other things being equal the dead weight of the running gear
per square inch of piston area increases as the scale of the
engine.
3. The twisting moments due to mechanical friction have
been neglected, as (so far as the present writer is aware)
the distribution and variation of the friction forces are not
known with any exactitude, and in any case one is a little
on the safe side in neglecting them. These friction moments
of course accumulate as one passes from the forward to the aft
end of the engine, where they amount in aggregate to about
15% of the mean indicated twisting moment, so their effect on
the forward end of the shaft is quite negligible.
90
DIESEL ENGINE DESIGN
4. The moment of inertia of the fly-wheel has been assumed
to be large in comparison with the fly-wheel effect of the
revolving and reciprocating parts of the running gear. In
cases where the fly-wheel is very small, or omitted altogether,
as in some two stroke marine engines, the irregularities of
turning effort are mainly absorbed by the angular acceleration
of the crank masses. Allowance is readily made for this effect
in the following manner. The combined twisting moment
curve for all the cylinders is first found without allowance for
fly-wheel effect, and a new zero line
is taken at the height corresponding
to the mean twisting moment. Or-
dinates measured to this new zero
line represent fluctuations of the
twisting moment from its mean
value. These ordinates are now
divided into segments proportional
to the fly-wheel effects of the fly-
wheel and crank masses. For ex-
ample, if there are four cylinders
and the moment of inertia of the
fly-wheel is three times that of one
set of crank masses, then the or-
dinates will be divided into seven
parts, one part being appUed in
opposite sense to corresponding
points on the twisting moment
curve of each cylinder and the re-
maining three parts of each ordinate
form the ordinates of a curve of
the twisting moments absorbed by
the angular inertia of the fly-wheel.
In the example worked out below
itgsdll be assumed that the fly-wheel
is large compared with the crank
masses, so that the process described
briefly above is not necessary.
Fig. 45 is a skeleton diagram of
the connecting rod positions for
every 20 degrees of revolution of
the crank-shaft for the determina-
tion of piston displacements.
CRANK-SHAFTS
91
Fig. 46 is a typical full load indicator card calibrated for
pressures vertically and percenta;ges of stroke horizontally.
Points corresponding to each 20 degrees of revolution are
marked on the diagram by scaling the piston displacements
ofE Fig. 45.
On Fig. 47 cylinder pressures are plotted on a crank angle
base from to 720 degrees (four stroke engine). The pressure
izo° m'miscf
during the suction and exhaust strokes is assumed atmospheric.
The inertia effect of the reciprocating parts per square inch of
piston area is plotted from the following figures : —
236 X 615
Inertia effect at top dead centre (1 +J) =5400 lb.
bottom „
236x615
236x615
g
(1-1) = 3600
90
g
X^= 9001b.
5
^^.^236X615^ l = 3200lb.
g ^2
in.)
Corresponding figures per sq. in. of piston area {78-5 sq.
are 69, 46, 11-5, and 41 lb. per sq. in. respectively.
The centrifugal forces of the revolving masses being radial
produce no twisting effect on the shaft. The dotted line
(Fig. 47) is the resultant of the pressure and inertia curves.
The twisting moments are now computed as in the following
table : —
92
DIESEL ENGINE DESIGN
toS
14,100
20,400
14,100
-11,200
-17,900
-15,100
-9,750
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CRANK-SHAFTS
93
The leverage tabulated in the second column is found by the
well-known graphical construction in Fig. 45, where the line of
the connecting rod is produced (if necessary) to meet the
horizontal line through the centre of the shaft, the intercept
being the leverage required to the same scale as the rest of
the diagram. The twisting moments are found by multiplying
the leverage in inches by the resultant forces per sq. in. of
•:^z^^
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94 DIESEL ENGINE DESIGN
piston area in lb. per sq. in. and by the piston area in sq. in.
(in this case 78-5 sq. in.).
Forces acting towards the crank are considered positive,
whether they are expansion forces or otherwise, and those
acting away from the crank negative. Assuming rotation
clockwise, leverages to the right hand of the centre line are
positive and those to the left negative. The signs of the
moments then look after themselves according to the signs of
their factors. It is not unusual to see the leverage in a case of
this sort treated as though it were always positive. The dis-
advantage of this proceeding is that in order to get the signs
of the moments correct those of the resultant forces have to
be reversed at every dead centre, which besides being incorrect
from a mathematical standpoint is inconvenient for the
draughtsman and confusing to others. The tabulated values
of the twisting moment are plotted in Fig. 48 (full line curve).
Identical curves for cylinders 3, 4, and 2 could be plotted in
their respective places at 180 degrees apart, in the order named,
but are omitted, for the sake of clearness. Dotted curve
numbered 2 is the resultant of the curves belonging to cylinders
1 and 2. Dotted curve 3 is the resultant of the curves belong-
ing to cylinders 1, 2, and 3, and so on. The simplest way of
obtaining these resultants is to trace the primary curve on a
piece of transparent paper and move it sideways into its required
position for the next cylinder, and then for every required
ordinate move the paper vertically (guided by the vertical
degree lines) until the zero line coincides with the top of the
ordinate of the curve to which it is required to add the effect
of another cylinder. In this position prick through the top of
the ordinate of the curve on the tracing-paper to the diagram
underneath. These resultant curves enable the twisting
moment at any crank-pin or journal at any angular position to
be read off the diagram.
Combined Effect of Ben^ng and Twisting. — It will be
seen that the peaks of the twisting moment curves occur about
30 degrees after the dead centres, and that the results previously
obtained for the bending moments with the cranks on dead
centre apply very closely to this position also, so that the
tabulated values of the bending moments at the various
journals and crank-pins combined with the twisting moments
existing at these points 30 degrees after the corresponding
firing dead centres have been passed, represent the maximum
CRANK-SHAFTS
95
conditions of stress at the points in question. The conditions
of bending when cranks 3 and 4 are on firing centre are of
course the same as those obtaining when cranks 2 and 1
respectively are in that position, the order in which the bending
moments occur being reversed.
For example, the bending moment at No. 2 crank -pin
when No. 4 cylinder is firing is the same as the bending moment
at No. 3 crank -pin when No. 1 cylinder is firing, and so on. A
comparison of the following table with the twisting moment
curves and the bending moments tabulated in the previous
articles will make the matter clear.
The equivalent twisting moment equals VT^+B^, and is
that twisting moment which would give the same shear stress
as the maximum shear stress due to the combined action of
twisting and bending moments actually obtaining.
Position.
Which
crank
30° past
firing
dead
centre.
Bending
moment
in. lb.
(see pre-
vious
tables).
Number
of twist-
ing
moment
curve.
Twisting
moment
in. lb.
from
curves.
Equivalent
twisting
moment
in. lb.
Maxi-
mum
shear
stress
lb. per
sq. in.
A. Journal
■ —
—
—
—
—
—
B. Crank-
pin
No. 1
No. 1
No. 2
No. 3
No. 4
105,100
16,100
79,000
24,400
—
—
105,100
3700
C. Journal
No. 1
No. 2
No. 3
No. 4
74,800
54,800
71,000
56,200
(1)
127,000
11,000
13,000
17,000
147,500
5200
D. Crank-
pin
No. 2
No. 1
No. 2
No. 3
No. 4
25,200
74,800
57,400
30,300
(1)
127,000
11,000
13,000
17,000
137,000
4820
E. Journal
No. 1
No. 2
No. 3
No. 4
62,600
80,600
80,600
62,600
(2)
112,000
115,000
31,000
31,000
140,000
4930
96
DIESEL ENGINE DESIGN
Which
crank
30° past
Bending
moment
in. lb.
Number
of twist-
Twisting
moment
in. lb.
from
Equivalent
twisting
Maxi-
mum
shear
Position.
firing
dead
(see pre-
vious
ing
moment
moment
in. lb.
stress
lb. per
centre.
tables).
curve.
curves.
sq. in.
F. Crank-
No. 1
30,300
112,000
pm
No. 2
57,400
(2)
115,000
128,500
4530
No. 3
No. 3
No. 4
74,800
25,200
31,000
31,000
No. 1
56,200
100,000
G. Journal
No. 2
No. 3
No. 4
71,000
54,800
74,800
(3)
93,000
93,000
45,000
117,000
4120
H. Crank-
No. 1
24,400
100,000
pm
No. 2
79,000
(3)
93,000
122,000
4300
No. 4
No. 3
No. 4
16,100
105,100
93,000
45,000
No. 1
82,000
2890
I. Journal
No. 2
No. 3
No. 4
(4)
82,000
82,000
82,000
Conclusions. — Maximum Shear Stress, 5200 lb. sq. in. —
Taking the fatigue stress in shear for mild steel, subject to
combined bending and twisting at 15,000 lb. per sq. in., the
factor of safety for a shaft newly lined up is about 3, and
diminishes very considerably as the bearings become worn out
of level.
The high values of the ggresses at the centre of the shaft
point to the advisability of making all couplings between
Sections of the crank-shaft of the full torsional strength of the
shaft, i.e. the aggregate shearing area of the coupling bolts
multiplied by the radius of their pitch circle should be equal to
the twisting modulus of the shaft. Thickness of coupling
flanges ^|^ diameter of the shaft.
CHAPTER VI
FLY-WHEELS
The Functions of a Fly-wheel are : —
1. To keep the degree of uniformity within specified
limits.
2. Where alternators running in parallel are in question
to limit the angular advance or retardation of rotation, to a
specified fraction of a degree ahead of or behind an imaginary
engine rotating with perfectly uniform angular speed.
3. To limit the momentary rise or faU in speed when full load
is suddenly thrown off or on.
4. To facilitate starting under compressed air.
In addition to the above the fly-wheel usually serves as a
barritig or turning wheel and a valve setting disc ; also the
inertia of the fly-wheel has great influence in determining the
critical speed at which torsional oscillations of the crank-shaft
are set up.
Fly-wheel Effect. — ^The fly-wheel effect of a rotating body
is its polar moment of inertia (mass X radius of gyration
squared) about its axis of rotation. For a fly-wheel or pulley
it is found approximately by multipljdng the weight of the rim
in pounds by the square of the distance in inches from the axis
to the centre of gravity of the section of the rim, the result
being in in.^ lb. units. This underestitnates the moment of
inertia slightly, and a more accurate method will be described
later. The fly-wheel effect of the running gear of one cylinder
is found with sufficient accuracy for most purposes by adding
the weight of the revolving parts (crank-pin plus unbalanced
part of two crank webs plus 0-65 of the connecting rod) to half
the weight of the reciprocating parts (0-35 of the connecting
rod plus cross-head plus piston-rod plus piston, etc.), and
multiplying the sum by the square of the crank radius.
For a screw propeller the radius of gjrration may be taken as
0-35 of the extreme radius if details are not available.
H 97
98 DIESEL ENGINE DESIGN
Degree of Uniformity. —
-p. r -i- -x Max. speed— Min. speed
Degree oi umformity = fj ^—^
° •' Mean speed.
Let d=Degree of uniformity.
Wi=Max. angular speed in radians per second.
W2=Min. „ „ „ „ „
Thend=2i^^i=:^ (1)
For a specified value of " d " the necessary fly-wheel effect
is calculated by means of the resultant twisting moment curve
of the engine. Let Fig. 49 represent the twisting moment
curve and the line A C the mean twisting moment. Let ABC
be the loop of largest area (with multi-
cylinder engines there are in general as
<y^ \C j^ many positive and negative loops in a
complete cycle as there are cylinders, and
Fio^ the area of each positive loop is the same
as that of each negative loop). If the
loop A B C is above the line A C, then the speed of the engine
is a minimum at A and a maximum at C, and the increase of
rotational energy of the fly-wheel, etc., between A and C is
equal to the work represented by the area of the loop ABC.
Let A = Area of loop A B C in sq. in. on the diagram.
E = Work represented by A B C in in. lb.
a = Scale to which turning moments are plotted in in. lb.
to the inch.
b= Scale to which crank-shaft degrees are plotted in
degrees to the inch.
A V a, V V)
Then E=— — - — , 57-3 being the number of degrees in a
radian.
Let WK2=Fly-wheel effect (moment of inertia) in in.^ lb.
_, , m WK^ w^
Then kmetic energy of wHeel= — -— ^ — (g=386 in. /sec. 2).
2 g
Change of kinetic energy from A to C
(Wi2-W22)=— — (Wi-Wa) (Wi+W2)= —— — (Wi+Wa)"
2g ^'^ ■■" 2g -'^ ■■---' ■■■" 4g
=WK2.d.w2(MEAN) -^g
if the difference between Wj and Wg is small.
But the change of kinetic energy is equal to E
PLY-WHEELS
99
and d =
Ex 386
or WK2=
Ex 386
-(2)
WK2.W2 "' "" w^.d
Example : Single cylinder engine 10" bore x 15" stroke.
Revs. 300. Turning moment diagram as in Fig. 48, full line.
E = 151,000 in. lb. Radius of gyration of wheel 30". Required
to find the weight of the wheel to give a degree of uniformity of
^/^O- 300x27r „^ ^ ,.
w= — ^:^ — =31'4 radians per sec.
W.K2=
60
Ex 386
151,000x80x386
W
=4,730,000
m.
d 31-42
Ply-wheel effect of running gear (267 H — — -) x 7-5^=21
'lb. \ 2 /
600
WK2 for fly-wheel=4,730,000-21,600 = 4,708,400 in.^ lb.
but K=30"
,. W = l^^^« = 5230lb.=2.34tons.
Twisting Moment Diagrams for two and four stroke engines
having from one to eight cylinders are shewn in Pigs. 50 to 61.
These have been drawn for an engine 10" bore by 15" stroke.
As the twisting moments of two engines of different sizes are
proportional to the bore ^x stroke, these curves may be used
for engines of any size by multiplying the moments by the
bore 2 (in inches 2) xthe stroke (in inches) and dividing by 1500.
The excess energy represented by the largest loop in each
diagram is given in the schedule below for each case.
PouE Stroke Engines. Two Stroke Engines.
Number of
Cylinders.
E in in. lb.
E in in. lb.
E in in. lb.
E in in. lb.
for ICxlS'
for l"xl"
for 10"xl5"
for rxl"
Cylinder.
Cylinder.
Cylinder.
Cylinder.
1
151,000
101-0
125,500
83-7
2
127,500
84-8
58,500
39-0
3
87,300
58-2
48,700
32-4
4
38,500
25-7
39,000
26-0
6
39,100
26-1
11,150
7-4
8
31,700
2M
2,200
1-5
100
DIESEL ENGINE DESIGN
Substituting those values of E for a cylinder 1 in. X 1 in. in
equation (2) the following formula is obtained : —
C.B2.S
WK2=
im^-
,100/
Where B=Bore of cylinder in inches.
S=Stroke in inches.
n= Revolutions per minute.
Values of C are given in the following schedule :-
No. of
Cylinders.
C for 4 Stboee Enoinb.
C for 2 SmiOKB Engini!.
1
355
243
2
298
137
3
204
114
4
90
91
6
91
26
8
74
5
Values for " d " used in Diesel Engine Practice. — For
certain purposes, as for instance spinning miUs, a fine degree of
uniformity is desirable, and d=about y^.
For direct coupled continuous current dynamos d=^ is
sufl&ciently fine to prevent flickering of lights, and may be used
unless considerations of momentary governing demand a
heavier wheel than the use of this figure would give rise to.
For marine engines and land drives, where regularity of
turning is not of importance, " d " may be about j^^.
The above values for the degree of uniformity must be used
with caution, as in a large number of cases (particularly four
stroke engines of six cylind^B and upwards and two stroke
engines of three cylinders and upwards) the considerations
discussed in the next article outweigh those of regularity in
turning.
Momentary Governing. — ^Under the head of governing it is
usually specified that the rise in speed when the load is thrown
off suddenly or the fall in speed when the load is suddenly
thrown on shall not exceed a certain percentage (usually
between 5 and 12) of the mean speed. Actually the governor
PLY-WHEELS
101
has relatively small control over this rise or fall of speed, as at
the instant when the load is thrown off suE&cient fuel has
already been deposited in the pulverisers to carry the engine
against full load for a period which may be anything up to
^ — V
r ^-^
- '\^ f '
- 4^ t
- 4^ A
- t A
r 7
_ x.t
_A~
T]~
^r
^
^.
N
>
- 'r.^
,
- ^^
^"
,x
V
- =^^
ID
^^
\
7
O
t
■
: 7
. y"
-^ it
-^t
' L
-Z-
_,Z _
[~
■
, >
o o
dj
H
'A
O
•A
o
A
H
CO
&
O
OOC\J — OCiOO'^^ »0<:jcn 0\J**O-- C\jC*3^uliO
two revolutions in the case of a four stroke engine. The brake
energy developed during this period is entirely devoted to
accelerating the fly-wheel and other rotating masses. Owing
to the fact that the governor does not act immediately the load
is thrown off, the wheel should be capable of absorbing the
102
DIESEL ENGINE DESIGN
/\ ""■ 1 1
L L 2 Cyli iders lO'^ /5 "
_i I
4 A
7 i ,
I — ^^5.
t ^
t V i
t _. 5
^ -- -V -- ■-
N
^ Z5
^^ V t-.
S Z^ iJ
0° m'' ak
Fig. 51.
3f0» ^0
-
1
~
1 1
4 (.yiinders Iff' 15
^
1
\
1
\
/
\
1
\
/
\
/--
...
-4
—
/-
--
—
--
J
J
ISO'
Fig. 52.
^ [-
A - -
4 - -
I ., .
f ._ .
t- = .-.
^
, .
r— 1
S Cwlinder. ;
k^lf
0° 30'
FlQ. 63.
/4
/3
/2
/;
w
9
8
7
e
6
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Fig. 55.
FOUR STROKE ENGINES.
PLY-WHEELS
103
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Fig. 61.
TWO STROKE ENGINES.
104 DIESEL ENGINE DESIGN
whole power of the engine for about three revolutions in the
case of a four stroke engine, about 1 -5 of a revolution in the case
of a two stroke engine.
Example : B.H.P. of engine (four stroke). 180
Revolutions per minute . . 375
Momentary rise in speed when
full load is suddenly thrown
off 12%
Radius of gyration of wheel . 18 in.
It is required to find the weight of the fly-wheel, neglecting
the fly-wheel effect of the running gear.
„- , , , ^. ^, ,,, ,180x33,000x12. „
Work done per revolution at full load r^^ in. lb.
Energy corresponding to three revolutions
180x33,000x36
375
=570,000 in. lb.
Angular speed at full load — — — =39-3 radians per second.
Momentary angular speed when load is suddenly thrown off
39-3 X 1 •12=44-0 radians per second.
If W= weight of wheel, then : —
Wx 182
•; „ ° {442-39-32) = 570,000 in.^ lb.
2x386 ^ '
and W = 3470 lb.
Alternators in Parallel. — ^When two or more alternator sets
are being run in parallel it is a necessary condition for working
that they keep almost exactly in phase. Due to inequalities
of twisting moment, slight differences of phase inevitably occur,
and these give rise to synchronising currents between the
various machines, the tendency of these currents being to
accelerate the lagging machinfp and retard the leading ones.
This effect keeps the whole system in a state of stability, but
cannot be relied on to correct any large fluctuations, and on
this account it is usual to specify that the maximum deviation
from uniform rotation shall not exceed three electrical degrees
on either side of the mean. If the alternator under considera-
tion has a field of two poles only, then the electrical degrees
correspond to crank-shaft degrees. In general, if the number
of pole pairs is " p," then one crank-shaft degree corresponds
FLY-WHEELS 105
to " p " electrical degrees. So far as the engine designer is
concerned, then, the problem consists in ascertaining the fly-
wheel effect required to keep the cycUc fluctuations on the
engine fly-wheel within a certain number of degrees, or more
commonly within a certain fraction of a degree of revolution,
oil either side of the mean.
The method of calculation may be described briefly thus :- —
(1) Assume any convenient figure for the fly-wheel effect,
e.g. 100,000 in. 2 lb.
(2) Plot twisting moment curve for complete period, taking
the zero of ordinates at the mean twisting moment.
(3) Reduce crank angles to time in seconds, assuming uniform
rotation.
(4) Reduce twisting moments to angular acceleration in
degrees per second^ by dividing by the assumed fly-
wheel effect and by the acceleration due to gravity
(386 in. per sec. 2) and multiplying by the number of
degrees in a radian (57-3).
(5) Plot angular acceleration to time or crank angle base.
(6) Integrate by planimeter, or otherwise, obtaining angular
speed curve.
(7) Integrate again, obtaining angular displacement curve.
(8) Measure maximum deviation from the mean position in
degrees.
(9) Increase or decrease the assumed fly-wheel effect in pro-
portion as the angular deviation so found is more or
less than the deviation specified. This gives the fly-
wheel effect required.
Example : Three cylinder, four stroke engine : —
Bore 20 in.
Stroke 32 „
Revolutions per minute . .150
Number of pole pairs ... 20
Angular deviation. 3 electrical deg.
Twisting moment curve as in Fig. 62.
The mean twisting moment being
taken as the basis.
The whole calculation is contained in the table below, in con-
junction with Figs. 62, 63 and 64.
Since the engine makes 150 revolutions per minute, therefore
20° = -|^=0-0222sec.
150x360
106
DIESEL ENGINE DESIGN
Assume fly-wheel effect of 10® in.* lb. for purposes of calcula-
tion. Then : —
Twisting moment x 386
. , , ^..,^. ^-. „^„^o T.M.
m degrees per sec.*=^ —
Acceleration in radians per sec.* —
T.M. X 57-3x386
10«
45-3
Referring to the table below : —
Values given in column 2 are scaled off Pig. 62.
„ „ 3 are obtained by dividing those
in column 2 by 45-3.
„ „ 4 are obtained by multiplying
the values in column 3 by
0-0222 sec.
Column 5 is obtained by successive addition of speed
increments.
Column 6 contains corrections necessitated by the fact that
the resultant of column 5 is not zero, owing to errors.
Column 7 gives corrected speeds which are plotted in Fig. 63.
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Fig. 64.
Columns 8, 9 and 10 are obtained similarly to columns 2,
4 and 6.
Total swing in phase (see Fig. 64) 24+28 = 52° or 26° each
side of the mean.
FLY-WHEELS
107
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108 DIESEL ENGINE DESIGN
Q
Allowable swing, 3 electrical deg. = —= crank-shaft deg.
JO
Fly-wheel effect assumed for calculation, 1,000,000 in.^ lb.
Therefore fly-wheel effect required
1,000,000x20x26 ,„„ ,„. . .„
=— — = 173xl0*in.2 lb.
If radius of gyration of wheel is 65 in.
Weight of wheel=gg^^^^ = 18-3 tons.
Allowance for the fly-wheel effect of the alternator rotor
would reduce this figure a little.
Torsional Oscillations and Critical Speeds. — In the great
majority of Diesel Engines the critical speed at which torsional
oscillations of the crank-shaft would occur lies far above the
practical range of the engine. Serious cases do arise occasion-
ally, however, and will arise more frequently when high speed
two stroke engines of six cylinders and upwards become more
common. The trouble when it arises is generally disposed of
very simply by altering the fly-wheel effect. Supposing for the
moment that a marine engine is under consideration. Then
the fly-wheel, the crank masses, the propeller and the shafting,
etc., constitute an elastic system having a natural frequency of
torsional oscillations which depends on the amounts and
positions of the fly-wheel effects of its component parts and on
the stiffness of the shafting. If this natural frequency happens
to be the same as the frequency of the torsional impulses due
to the working strokes of the engine, then the oscillations tend
to become accumulative, vibrations are felt in the shafting,
and the crank-shaft may hammer in its bearings. Then the
engine is running at a critical speed. In general a two stroke
engine gives as many torsional impulses per revolution as there
are cylinders and a four stroke engine half this number, so that
a two stroke engine attains i^s critical speed at one-half the
number of revolutions required by a four stroke engine. For
example : Suppose the crank-shaft, etc., of a six cylinder four
stroke engine has a natural frequency of 2400 complete oscilla-
tions per minute, then the critical speed will be 800 revolutions
per minute. A similar engine working on the two stroke cycle
would have a critical speed of 400 revolutions per minute. It
will be obvious that all the revolving masses in connection with
the crank-shaft (apart from trifling items) must be taken into
FLY-WHEELS
109
consideration, so that the critical speed of a marine engine
coupled to a dynamo for testing purposes will be different to
that obtained when the cBigine is installed in the ship.
Natural Frequency of Torsional Oscillation. — Consider the
simple system shewn in Fig. 65,
consisting of a shaft fixed at
one end and carrying a fly-
wheel at the other. If the fly-
wheel be turned through an
angle against the torsional re-
sistance of the shaft and then
released suddenly the system
will oscillate until the energy
has been dissipated in friction,
the angle through which any
section of the shaft oscillates
distance from the fixed end
rm^mai
.^
L
1
u^^w
Mk^
Fig. 65.
being proportional to the
The latter is called the Node.
Let f =Frequency in complete oscillations per second.
F=Frequency „ ., „ minute.
d=Diameter of the shaft in inches.
I=Polar moment of inertia of shaft section in inches*.
Z= Length of shaft in inches.
W = Weight of the wheel in lb.
K=Radius of gyration of the wheel in inches,
g = Acceleration due to gravity in inches per second^ (386)
G=Modulus of rigidity of the shaft material (about
12,000,000 lb. per sq. in.).
Then :—
f
=^Vj^>-"=-V,^--'"
If the shaft is not of the same diameter throughout its length,
but consists of sections of length ?i, h, etc., of diameter dj, dj,
etc., then the equivalent length ^o of shift of standard diameter
do is given by
''<^Mf)'^- ■ ■ *■
For example : 1 ft. of 6 in. shafting is equivalent to 16 ft. of
12 in. shafting, 16 ft. being ( -g ) • For this reason, the
lengths occupied by coupling flanges, etc., are negligible.
110
DIESEL ENGINE DESIGN
If there are a number of fly-wheels or other rotating masses
at different distances from the node, then the frequency is
given very closely by the following : —
^-'•''^m^
etc.
-(2)
In the case of an engine and shafting no point on the latter
is fixed, and the position of the node is to be inferred from
considerations of dynamic equilibrium. It will be obvious that
a uniform shaft with an equal wheel at each end will have its
node at the centre, the oscillations of the two wheels being
equal in magnitude and opposite in sense at every instant.
It is almost equally apparent that in a similar case, with un-
equal wheels the position of the node will divide the shaft into
two lengths in inverse ratio to the fly-wheel effects at their
ends.
In general, the position of the node is determined with
sufficient accuracy to enable the critical speed to be predicted
within a few revolutions per minute by treating the fiy-wheel
effects (moments of inertia) as though they were weights and
locating the node at their centre of gravity.
A crank-shaft may be treated as a uniform shaft of the
diameter of the journals.
Example : Six cylinder, two stroke marine engine.
Bore 24 in.
Stroke ....
Revolutions per minute
Diameter of shaft
Lengths of crank-shaft
Weight of fly-wheel
Radius of gyration of fly-wheel .
Weight of revolving parts for one crank
,, reciprocating* arts .
,, propeller ....
Radius of gyration of propeller .
Shafting
. 35 in.
120
. 15 in.
as in Fig. 66
12,000 lb.
. 40 in.
4,500 lb.
51,000 lb.
10,000 lb.
. 20 in.
as in Fig. 66
Reduction of Shafting to Standard Diameter of 15 Inches.
Length of 14" shafting 620 in.
Equivalent length of 15" shafting, 62o(ijy =815 in.
FLY-WHEELS
111
Equivalent length of 15* shafting between the pro-
peller and fly-wheel = 150 + 81 5 + 60 = 1025 in.
Moment of inertia (polar) of l&"shaft=^x 154=49,600 in.*
[_^:^
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112 DIESEL ENGINE DESIGN
Fly-wheel Effects.—
W.K2 for propeller = 10,000 X 20^=4,000,000 in. lb.
W.K2 for crank masses =(4500 +2500) 6 x 17-52
= 12,900,000 in.Mb.
W.K2 for fly-wheel = 12,000x402 = 19,200,000 in." lb.
Position of Node. — Take moments of fly-wheel effects
about A : —
19-20x1025 = 19,700
12-9 (1025+37 + 48 + 74+24) =15,600
Total 35,300
The distance of the node from A=— ^-p=978 in.
Dealing with the part of the system to the left of the node
and applying equation (1) : —
t:, „ -- 749,600 X 12 X10«X 386 „„„. .,, ,.
P=9-55 . / — '—- i =2320 oscillations
V 978x4xl0« per minute.
2320
And the critical speed — -- =386 R.P.M.
which is far above the working range of the engine.
Critical speeds of the second and third order, and so on, are
possible at 193, 128, 96 R.P.M., etc. ; but these are not, as a
rule, important.
It is sometimes useful to repeat the calculation for different
weights of fly-wheel and plot a curve connecting moment of
inertia of fly-wheel and critical speed, so that the possible
variation of critical speed obtainable by altering the fly-wheel
can be seen at a glance.
Torsional oscillations about two or more nodes are also
possible, but as these involve higher speeds, and are therefore
more subject to damping, it is doubtful if they are of much
practical importance. «
To find the Moment of Inertia of a Fly-wheel. — In the first
instance, suppose the wheel in question is a disc wheel, i.e. a
solid of revolution. Referring to Fig. 67, the thick, full line
represents the section of the wheel. Z Z is the axis and S S is a
line through the extreme radius of the wheel parallel to the
axis at a distance R from the latter. Rule any line A B parallel
to the axis, cutting the outline of the section in A and B.
Project A and B on to S S at C and D. Join C and D to any
FLY-WHEELS
113
convenient point on the axis, cut-
ting A B in Aj and Bj. Proceed
similarly with different positions of
the line A B and join up the various
positions of A^ and B^, thus obtain-
ing a new figure — the First Derived
Figure. Treat this figure as though
it were the original figure, and ob-
tain the Second Derived Figure.
Similarly with this figure obtaining
the Third Derived Figure.
9
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D
S
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Fig. 67.
Let A = Area of original section in sq. in.
A 1== Area of First Derived Figure in sq. in.
A2=Area of Second Derived Figure in sq. in.
A3=Area of Third Derived Figure in sq. in.
w= Weight in lb. of one cubic inch of the material.
(■E 1 P 1 p
ThenA= y.dx, Ai=g^j x.y.dx, Aa^^J x^.y.dx,
R3J0
^.y.dx.
Weight of wheel=2w.w 1 x.y.dx=27r.w.R.Ai
Moment of inertia of wheel=27rw I x'y.dx=27rw R^A3
Radius of gyration ''=R^.-^
Ai
The above hold good for any position of the line S S, which
may therefore be taken where most convenient. In cases
where the section tapers towards the extreme radius (a screw
propeller, for instance) the line S S is best located at a distance
of about one-half or one-third of the extreme radius from the
axis.
In the case of a screw propeller or a fly-wheel with arms, the
rotating body must first be reduced to an equivalent disc wheel.
This is readily done as follows : Describe a radius R which cuts
through the arms or blades, as the case may be. Divide the
total area of section at this radius by 27r.R and the result is the
thickness of the equivalent disc at this radius. Repeat for a
number of different radii covering the whole range.
Types of Fly-wheels. — ^Fig. 68 shews a disc wheel cast in one
114
DIESEL ENGINE DESIGN
piece and provided, with a number of drilled holes in the rim for
turning the engine by means of a bar. Degree marks are cut
on the edge of the rim to facilitate valve setting. Fig. 69 shews
a disc wheel with a separate centre, an arrangement which
Fm. 68.
Via. 69.
makes it easier to obtain a sound casting. Large wheels are
usually cast in two pieces, and Fig. 70 shews a design which is
suitable for weights up to at least 20 tons. It should be noted
that no keys are provided for securing the wheel to the shaft.
If the boss of the wheel is bored one-thousandth per inch of
Fig. 70.
diameter less than the shaft and the bolts are drawn up at
about the temperature of boiling water the frictional grip is
quite sufficient for the largest wheels and the danger of split-
ting the boss of the wheel involved in the use of keys is avoided.
The same applies to pulleys for belt or rope drives. A rather
more elaborate wheel, in which greater precautions have been
FLY-WHEELS
115
taken, is shewn in Fig. 71. In this case it is advisable to make
the bore of the wheel the same as the shaft diameter and to give
a shrinking allowance of about one -thousandth per inch of
diameter to the bore of the shrunk ring. For large stationary
engines some form of barring gear is necessary, and where
electric power is available a motor-driven gear is a great con-
venience. For marine engines a worm, or other self -locking
gear, is essential, and where the auxiliaries are electrically
driven an electric turning gear should be fitted, as the use of
the latter greatly expedites adjustments to the valve gear.
Fig. 71.
Strength of Fly-wheels. — Continental practice favours a
peripheral speed of about 100 feet per second for cast iron fly-
wheels, and this corresponds to a stress of about 1000 lb. per
sq. in. An investigation by Mr. P. H. Smith into the case of a
split fly-wheel which burst at Maidenhead in 1912 shewed that
the engine (the fly-wheel of which had a normal working peri-
pheral speed of about 100 feet per second) was running about
double its normal speed, and as the stress varies as the square
of the speed, it follows that the factor of safety under normal
conditions was about 4. Destruction tests of wheels and
models of wheels shew that the bursting speed is about 200 feet
116 DIESEL ENGINE DESIGN
per second for split wheels and 400 feet per second for solid
wheels. The discrepancy seems very large and difficult to
account for. Average British practice is in favour of a slightly
lower peripheral speed (about 90 feet per second). With
marine engines considerations of space generally necessitate a
still lower figure. The strength calculations for a fly-wheel
will be illustrated by an example.
Example : Required to find the approximate dimensions
of a fly-wheel suitable for 1 80 revolutions per minute given that
W.K2=40,000,000 in.2 lb.
Peripheral speed, 100 ft. per sec.
Maximum twisting moment due to engine, 450,000 in. lb.
Outside radius of wheel = — — — — =63-7 in, say 64 in.
27rXl80 •'
Take inside radius of rim =52 in.
Then radius of C.G. of rim section=58 in.
Let B= width of rim. Then: —
Weight of rim = I2xBx2irx 58x0-26
And approximate moment of inertia
= 12 xB X 1-64 X 583=40,000,000 in.^ /lb.
From which B = 10-5 in.
Since the stress due to a peripheral speed of 100 ft. per sec.
is about 1000 Ibs./in.^, the total tension at each joint of the
rim is equal to 12x10-5x1000 = 126,000 lb., for which pull
the dowel and cotter section must be designed.
Allowable stress in dowel, say 6000 lb. per sq. in.
Effective area of dowel section ^ ' ^ =21 sq. in.
6000 ^
Since about one-third of the section of the dowel is cut away
by the cotter hole (see Pig. 72), the gross sectional area of the
21 X 3
dowel must be — - — =31-5 sq. in., say 4 in. X 8 in.
8 •
Thickness of cotter = — "about 2| in.
Bearing pressure of cotter on dowel
126,000 „.-„,,
2-75x4 ^ ' P®^ ^^' ^^'
which is allowable.
If the hole for the dowel is made i in. wider than the dowel
itself, the bearing length for the cotter on the rim will be
10-5 — 4-5 = 6 in., and bearing pressure of cotter on rim
FLY-WHEELS
117
126,000
=7650 lb. per sq. in., also allowable.
2-75x6
Allowable shear stress for cotter (which is in double shear),
say 5000 lb. per sq. in.
Depth of cotter ' =9-2 in., say 9| in. over the
rounded ends. 5000x^-75
The distance " 1 " between the inside edge of the cotter hole
and the rim joint must be sufficient to obviate risk of the
Fig. 72.
intervening metal being torn out in double shear. (This point
is sometimes overlooked in otherwise well-proportioned rim
joints.)
Allowing a shear stress of 1000 lb. per sq. in. : —
, 126,000 ,. _ .
1= ,^^^ ^ — - = 10'5 in.
1000 X 2 X 6
118 DIESEL ENGINE DESIGN
Owing to the difficulty of analysing the straining actions on
the arms it is well to give the latter ample proportions.
An approximate method of calculation is given below.
Assume that the maximum twisting moment due to the
engine (450,000 in. lb.) is transmitted to the rim by means of
a constant shear force across the arms, and that the be'nding
moment is a maximum at each end of an arm and zero at the
centre.
The length of each arm from boss to rims is about 40 in.,
and the distance of its centre from the centre of the wheel
about 32 in.
Then shear force in each arm= ' =2340 lb. (assuming
six arms). ^^^^
And maximum bending moment at end of each arm =
2340x20=46,800 in. lb.
Taking a low stress of 500 per sq. in., to allow for direct
tension in the arms, bending modulus- of arm section
500
This is satisfied by a rectangle section 6 in. x 10 in., which could
be replaced by an oval section about 7 in. x 12 in. to reduce wind
resistance. The bolts at the hub of the wheel are sometimes
made as strong as the rim joint, in which case the core area of
two bolts will be the same as the net efEective area of one dowel,
viz., 21 sq. in.
This gives a bolt of about 4 in. diameter. If shrunk rings are
employed these will have a square section about 10-5=(3iin.)^-
The above calculations must be regarded as preliminary only,
and give the draughtsman a basis on which to start designing.
The next step wiU be to check over the weight and radius of
gyration of the complete wheel in the manner already describe r* .
The dimensions and stress calculations will then be amende
accordingly. ^
Literature. — For information on the strength of fly-wheel ,
see : — Unwin, W. C, and Mellanby, A. L., " The Elements of
Machine Design," Part II.
For information on critical speeds, see:— Morley, A.,
"Critical Speeds for Torsional and Longitudinal Vibrations,"
Engineering, December 9th, 1910.
CHAPTER VII
FRAMEWOBK
A LARGE number of diflEerent types of framework have been
employed in Diesel Engine construction, and a complete
classification will not be attempted here. The outstanding
types in successful practice may, however, be broadly divided
into a few well-defined classes, as under : —
/
«' A " Frame Type. — This is the earliest type of Diesel
Engine construction, and on account of its merits is stUl very
extensively used. Referring to the diagrammatic drawing
Fig. 73, it wiU be seen that a stiff bedplate of box section is
provided, and that each cylinder stands on its own legs without
support from its neighbours. The legs of the column are cast
integrally with the cylinder jacket, into which a liner is fitted.
The breech end of the cylinder is closed by means of a deep
cylinder cover of box section.
119
120
DIESEL ENGINE DESIGN
The main tensile load due to the cylinder pressure is trans-
mitted from the cover through the jacket
and legs to the bedplate. The reaction
corresponding to this load occurs, of
course, at the main bearings, and con-
sequently that part of the bedplate be-
tween the column feet and the main
bearing housings must be designed to
deal with the bending moment occa-
sioned by the fact that the tensile load
in the columns and the reaction at the
bearings are not in the same plane.
Casting the cylinder jacket and column
in one piece reduces fitting and machin-
ing operations to a minimum and the
independence of the individual cylinders
would appear to have no disadvantages
so far as land engines are concerned.
Fig. 74 shews the same type of con-
struction applied to a two stroke land engine and Fig. 77 to
a four stroke marine cylinder.
Fig. 74.
Fi8. 77.
Fio, 78.
FRAMEWORK
121
Slightly lengthening the column legs enables crosshead and
guides to be fitted (see Fig. 78, which represents a large two
cycle land engine). Occasionally one of the column legs takes
the form of a steel tie rod, with a view to giving greater accessi-
bility to the running gear and to enable the crank-shaft to be
replaced, if necessary, without dismantling the whole engine.
Unfortunately this arrangement nullifies many advantages of
the " A " frame construction, as special splash guards must
now be fitted to retain the lubricating oil, which office they do
not always perform very efficiently, and also additional
Fig. 75.
Fio. 76.
machining and fitting operations are introduced which add to
the cost of production without increasing the efficiency of
working. Figs. 75 and 76 shew this construction applied to
trunk and crosshead engines respectively.
Crank-case Type. — The crank-case type of Diesel Engine
was introduced when a desire was felt for higher speeds,
necessitating forced lubrication. The crank-case bears external
resemblance to that of a high speed steam engine (see Fig. 79),
On the other hand, the high pressures dealt with in the cylinder
of a Diesel Engine necessitate the crank-case being strengthened
internally to an extent which is not found necessary in steam
practice. Sometimes the box or girder construction of the
crank-case is relied upon to transmit the tensile stresses from
122
DIESEL ENGINE DESIGN
the cover to the bedplate ; more frequently, however, steel
staybolts are provided for this purpose (see Fig. 80). The
latter procedure, however, does not justify flimsy construction
Fig. 79.
or a careless distribution of metal in the crank-case, as the
guide pressure has still to be reckoned with and the pressure
caused by tightening up the staybolts may be relied upon to *
cause serious distortion of a poorly ribbed case.
Fig. 80.
Fig. 81.
The cylinders being separate are secured either by a round,
studded flange or by passing the staybolts through each corner
of a deep, square flange of hollow section cast at the lower end
FRAMEWORK
123
of the cylinder jacket for this purpose. The latter arrangement
requires four staybolts for each cylinder, whereas with the
former it is usual to arrange a pair of staybolts only at each
main bearing girder. With the crank-case construction it is
not necessary to make the side girders of the bedplate so strong
as for an " A " frame type of engine, as the bending action
referred to above is avoided and the bedplate and crank-case
when bolted together form a girder construction of great
rigidity. On the other hand, the upper part of the crank-case
is clearly subject to bending actions similar to those which
^t
^d
'^
Fig. 82.
Fis. 83.
occur in the bedplate of an " A " frame engine and must be
designed with this fact in view. Fig. 81 shews a section through
a two stroke trunk engine of the crank-case type. Figs. 82
and 83 shew the crank-case construction applied to crosshead
engines. The suitability of this type of framework for marine
service has been amply proved in practice. In some cases the
crank-case is common to two or more cylinders, and in others
the case for each cylinder is a separate casting, the individual
cases being bolted together to form a virtually continuous box
of great strength and rigidity.
Trestle Type. — With this construction, illustrated in Fig. 84,
124
DIESEL ENGINE DESIGN
the cylinders are bolted to a base plate or entablature resting
on trestle-shaped columns, the feet of the latter being secured
to the main bearing girders. If the guide casting itself and its
attachment to the trestles are sufficiently strong, this construc-
tion can compete with the crank-case type of frame in the
matter of rigidity. The same effect could doubtless be achieved
by some form of bracing. In Fig. 84 an alternative form of
crosshead and guide is shewn, to which the trestle arrangement
M
7^
E^l
r
Fio. 84.
y I H
particularly lends itself, viz.,lfche fore and aft double guide
usually found on paddle steamers. The advantages of this
form of guide for Diesel Engines are : —
1. Accessibility of running gear, the piston cooling gear in
particular.
2. The guide blocks being free to adjust themselves to the
guides are capable of sustaining a greater specific
pressure, and consequently require less bearing surface
than the usual type of guide shoe.
FRAMEWORK
125
It is to be noted, that the trestle construction involves some
little extra care to retain lubricating oil when forced lubrica-
tion is used.
Staybolt Construction. — ^With this construction the cylinders
are connected to the bedplate by means of turned bolts only,
and the saving in weight on this account amounts to the
considerable figure of about 25% of the complete weight of
the engine. The reduction in cost of manufacture must also
be considerable when the staybars are made of ordinary bright
shafting screwed at the ends. Examples are shewn in Figs. 85
Fig. 85.
Fio. 86.
and 86 for trunk and crosshead engines respectively. In the
latter the cast columns are relatively light, being designed for
the guide pressure only, and their top ends are made free to
slide, in order to avoid the subjection of the column to tensile
strains on the compression and firing strokes. The difficulty
of enclosing such engines adequately renders them unsuitable
for forced lubrication under circumstances where economy of
lubricating oil is a consideration ; on the other hand, large
slow-running engines, fitted with drip or ring lubrication,
appear to be quite satisfactory when built on this principle.
Design of Bedplates. — ^The design of a suitable bedplate
involves consideration of the following points, which will be
dealt with in order, viz. : —
126 DIESEL ENGINE DESIGN
(1) The provision of a suitable main bearing.
(2) A girder construction under each main bearing, capable
of supporting the full bearing load without central
support.
(3) A sufficiently strong and stifE connection between the
main bearing girders, forming at the same time an oil-
tight tray.
(4) Suitable studding or staybolt arrangements for carrying
the tensUe pull of the columns.
(5) Arrangements for supporting the cam-shaft driving gear.
(6) Means for collecting drainage of lubricating oil to some
convenient sump, whence it can readily be drawn off
With a view to filtration and repeated use.
(7) Facings for barring gear, auxiliary pumps, etc.
Main Bearings. — Examination of badly worn crank-shafts
indicates that the high bearing pressure obtaining for a short
time when the piston is at its firing centre gives rise to far less
abrasive action on the bearings than the less intense but longer
sustained pressures due to inertia and centrifugal force in a
four cycle engine. It appears that a film of oil is capable of
sustaining a heavy pressure for a short time, but once the film
has broken down, relatively feeble pressure is sufficient to
cause abrasion, and there is small chance of the surfaces
receiving a new film of lubricant until the pressure is removed.
The result is that very little trouble is experienced with the
lubrication of the main bearings of four stroke engines (the
pressure on the journals being frequently reversed) unless the
peripheral speed is so high as to reduce seriously the viscosity
of the oil film by means of the heat generated by friction. In
this case forced lubrication improves matters up to a certain
point, but if the speed is still further augmented there comes
a point when the bearings requite to be water cooled.
With two cycle engines the direction of pressure is probably
not reversed at all in most cases, and lubrication is con-
sequently more difficult. When the peripheral speed is low,
and the oil film in consequence as stable as possible, satis-
factory results are obtainable even with ring lubrication of
good design, if the maximum bearing pressure is kept about
30% lower than would be considered good ordinary practice
with four stroke engines. If, on the other hand, high peripheral
FRAMEWORK
127
speeds, or moderate bearing surfaces, or both, are required,
then a system of high pressure forced lubrication would appear
to be necessary, preferably in conjunction with a system of
water cooling in extreme cases. The following table gives a
rough guide to the limitations of the various systems of main
bearing lubrication : —
System of Lubrication,
Peripheral Speed
of Journal,
feet per minute.
Projected Area
of One Journal
expressed as
percentage of
the Area of
Piston.
FoxjE Stroke Engines —
Ring lubrication
Forced lubrication .
550
750
55%
40%
Forced lubrication and water
cooling
above 750
40%
Two Stroke Engines —
Ring lubrication
550
75%
Forced lubrication .
700
60%
Forced lubrication and water
cooling
above 700
60%
Drip or Syphon Lubricated Bearings are not used on land
engines, but are sometimes fitted to Marine Diesel Engines of
the open type. Apart from the fact that the caps have only to
be proportioned to the inertia and centrifugal loads, these
bearings are similar to those provided for steam engines, and
need not be described here.
Ring Lubricated Main Bearings. — ^These are similar to the
bearings fitted to electrical machinery and need not be described
in detail. The arrangements for catching the oil squeezed out
of the bearings and conveying it back to the oil well merit
careful attention, as inefficiency in this direction leads to un-
necessary waste of oil. In particular, the oil spaces and holes
should be as large as possible, to avoid congestion. Fig. 87
shews a very usual form.
Forced Lubricated Bearings. — ^These follow high speed
steam engine practice very closely, and the usual form of
128
DIESEL ENGINE DESIGN
staggering the circumferential oil groove in the top and bottom
brasses is generally adopted. See Fig. 88.
Main Bearings Shells. — These are of cast iron in commercial
work, and in the best practice
the shells are tinned previous
to the white metal being
poured in. The chief essen-
tials are : —
(1) Adequate thickness of
sheU.
(2) Good quality of white
metal.
(3) Good adhesion between
white metal and shell.
Common proportions are
shewn in Figs. 87 and 88, the
unit being the diameter of
the journal.
The bearing cap should be
designed as a beam capable of
carrying a central load equivalent to the full inertia and centri-
fugal load due to one set of running gear. This is possibly a
Cast Iron 0-23
Cast Steel '15
Fig. 88.
0-125-
little on the safe side, but reference to Chapter V will shew that
the margin is not large in the case investigated there.
Main Bearing Girder. — ^Where forced lubrication is used,
FRAMEWORK
129
the main bearing girder may conveniently be of I section, the
bottom flange being formed by the oil tray ; for ring lubrication
a box section lends itself more conveniently to the formation
of the oil reservoir. In any case the box section is preferable
in the larger sizes. The depth of the girder is determined by
that of the oil tray required to give an inch clearance or so to
the connecting rod big end at the bottom of its path. Referring
to Chapter V, it will be seen that the maximum reaction at a
bearing for the case considered is equal to 0-8 of the resultant
load due to pressure, inertia
and centrifugal force, and this
is the load for which the girder
must be designed. In other
cases the load may be less than
this, but it is doubtful if in any
case it approximates to the
conventional load frequently
assumed, viz., one-half the re-
sultant cylinder load.
A very debatable point is the
extent to which the oil tray
can legitimately be regarded as
a part of the tensile flange of
the girder. The author's prac-
tice in this respect is to ignore
the middle half of that part of
the tray lying between two
bearing girders (see Fig. 89).
The span of the girder is the
distance between the two points
at which it meets the side
girders.
If W =Load on girder in lb .
1 =Span in in.
M = Bending moment in
in. lb.
Then M== 0-2 Wl — approxi-
mately.
The assumption being that the
fixing moments at the ends are
negligible (which if not correct
130
DIESEL ENGINE DESIGN
is on the safe side) and that the load is distributed over the
journal. Allowable stress 1500-2500 lb. per sq. in. for cast iron.
Side Girders. — ^With " A " frame engines the bending
moment on each side girder may be taken as : —
Half pressure load x Distance between centres of bearings
6
The usual stress allowance being about 1500 lb. per sq. in.
Where the trestle or crank-case type of frame is used the side
girders may be of lighter section (proportions will be given
later).
Arrangements for Carrying Tensile Pull of Columns. —
With the " A " frame construction the foot of each column is
secured by a row of st^ds, the stress in which when referred to
the normal maximum working pressure of 500 lb. per sq. in. in
the cylinder amounts to about 5000 to 10,000 lb. per sq. in.,
according to the size of the stud. It is very convenient to have
a list of the loads which studs and bolts of different sizes can
conveniently carry, and such a list is given below : —
Size of Bolt
Stress (Core)
'
or Stud
allowed,
Working Load,
(Whitworth).
lb./m.=>
lbs.
i"
2000
240
r
2850
550
i"
3550
1080
i"
' 4250
1800
1"
5000
2750
H"
5250
3650
n"
5500
5000
If"
6000
6300
ir
7100
9300
If"
8500
15,000
2"
9tOO
21,500
2i"
10,000
28,000
2^"
10,000
37,000
2i"
10,000
44,000
3"
10,000
54,500
Care must be taken that none of the studs are at any consider-
able distance from adequate supporting ribs. This is best
FRAMEWORK
131
obtained by judicious spacing of the studs rather than the
provision of special ribs for the purpose.
With the crank-case and trestle t3rpes staybolts are usually
fitted, and in land work at any rate these should terminate
within the bedplate and not penetrate to the under side of the
latter for fear of oil leakage, which would- destroy the con-
crete. The studs or bolts used to secure the crank-case to the
Fig. 90.
bedplate may be disposed more with a view to making an oil-
tight joint than to carry any definite load. If staybolts are not
fitted, then a sufficiency of effective bolt or stud area must be
arranged in the neighbourhood of each column foot, and some
of the bolts or studs must be inside the crank-case.
Cam-shaft Driving Gear. — The motion required by the
valve ge£tr is derived from the crank-shaft by spiral or spur
gearing in the majority of designs. Fig. 90 shews a very
common arrangement of spiral drive, with the driving-wheel
between the two sections of a divided main bearing. It is good
132
DIESEL ENGINE DESIGN
practice to make the combined length of the two sections about
50% greater than the length of a normal bearing. There would
appear to be nothing against having the spiral wheel outside
the bearing altogether, provided the gear is at the fly-wheel
end. This position for the valve gear drive is preferable to the
compressor end as the weight of the fly-wheel tends to keep
the journal in contact with its lower bearing shell, whereas the
forward journal has freedom of motion (under the influence of
forces which vary in direction) to the extent of the running
clearance.
In six cylinder engines the spiral gear is frequently arranged
at the centre of the engine, where it is very easily accommo-
dated. There seems to be some feeling that the cam-shaft
would whip unduly if driven
^ from the end. This dif&-
■yy culty (if any difficulty can
^3 be said to exist) is easily
overcome by making the •
cam-shaft about 10% larger
in diameter than would be
considered sufficient for a
four cylinder engine.
Where spur gearing is
used for the valve gear
drive, facings must be pro-
vided for the support of
the first motion shaft.
Oil Drainage. — With
land engines of the non-
forced lubricated type the
oil which drips down from
the cylinders and is thrown
from the big ends is drained
periodically from the for-
ward end of the bedplate
and holes are cored through
the main bearings girders
to give the oil free passage.
Perhaps the best arrange-
ment is a rectangular duct
about four inches square
FiQ. 91. running down the centre of
FRAMEWORK 133
the oil tray. Small holes are useless as they are easily choked.
With forced lubricated engines the same arrangements are
made with the addition of a collecting sump of good capacity,
a pump for forcing the oil into the bearings and filters
in duplicate. These features being familiar in steam engine
practice need not be described in detail. It must be borne in
mind, however, that where trunk engines are being considered
the oil is contaminated with carbon, so that the filtering arrange-
ments require to be on a more liberal scale than is necessary with
engines in which the cylinder is isolated from the crank-case.
Proportions of Bedplate Sections. — ^Fig. 91 gives approxi-
mate proportions for various types of bedplate sections, the
unit being the stroke of the engine. Type "A" is usually
associated with the " A " frame construction. Type " B " is
a useful one for main or auxiliary marine engines as it enables
the engine to be bolted direct to a tank top or to a deck with-
out buHding up a special seating.
Type " C " is preferable to type " A " for land generating
sets as the extra depth of bedplate enables the generator to be
flush with the engine room floor without the necessity of
building the engine on an unsightly plinth. A deep bedplate
is also very desirable with six cylinder engines as the cancella-
tion at the centre of the engine of the inertia and centrifugal
couples gives rise to vibrations, the amplitudes of which are
reduced by increasing the stiffness of the framework.
The general thickness of metal may be about 4% of the
cylinder bore increased to about 6% or 8% on machined
surfaces. These figures are usually exceeded oh small engines
on account of the difficulty of obtaining consistent results in
the foundry with thin metal. The following figures for
different diameters of cylinder represent good practice : —
Bore of
General thickness of
Cylinder, in.
Metal for Bedplate, in.
10
f"
12
i"
15
if"
18
W
21
iiV
24
lA"
27
ir
30
If"
134 DIESEL ENGINE DESIGN
The above are useful as a guide, but must not be relied upon
without check calculations, as special constructions may-
require local strengthening to keep the stresses within a safe
figure.
"A" Frames. — ^An "A" frame for a four stroke trunk
engine is shewn in Fig. 92 ; the discussion of those parts which
are common to separate cylinders as distinguished from
cylinders combined with columns will be reserved for a
separate chapter. The liner is a push fit or light shrink fit in
the upper flange of the jacket. The fit at M should be a few
thousandths slack, to prevent seizure at this point. The jacket
is swelled locally, to give adequate water passage, and six or
eight strong ribs are provided to transmit the tensile load
across this section, which would otherwise be greatly weakened
by the discontinuity of the wall. At Q the liner is a push fit,
allowing the liner to expand axially relatively to the jacket.
Sometimes a stuffing-box is fitted to prevent water leakage.
P is the water inlet connection. L L are bosses to accommodate
lubricator fittings. N is a cleaning door.
Strength of " A " Frames. — Fig. 92 is drawn for a 15 in.
cylinder, the stroke being 21 in., and the dimensions given re-
present good average practice. The stresses may be checked
as follows : —
Maximum working pressure . . . 500 lb. per sq. in.
„ load 0-784x152x500
= 88,000 lb.
Tensile stress in parallel part of jacket (mean dia.=23-5")
88,000 no^,, r ,
= TT^n^ — ; =1190 lb. /m.^
:rX 23-5x1 '
Next consider section " AA " of one leg. For this section
conditions are worst if the nuts at the foot are not tight and
the reaction at the foot consists of a simple vertical pull of
44,000 lb. Referring to Fi^ 92, the direct tensile pull in the
leg is 42,000 lb., and th* sectional area at "AA" being
57 sq. inches,
Direct tensile stress at "AA" = 42,000 -^ 57 = 737 lb. per sq. in.
Shear stress „ =13,500-^57 = 237 „ „ „
Bending moment „ =44,000 x 9-5=42,000 in. lb.
Section modulus ,, =<^230in.^
Bending stress „ =420,000 ^230 = 1825 lb. per sq. in.
Total tensile stress „ = 1825 -|-737 =2562 „ „ „
FRAMEWORK
135
This stress refers to the outside fibres of the column, and is
probably in excess of what actually occurs as the fixation of the
foot by the holding down studs produces a moment in the
reverse direction.
Fro. 92.
136 DIESEL ENGINE DESIGN
Section " BB " is subject to maximum stressing action if the
foot is securely fixed to the bedplate, as it should be. The
vertical reaction of 44,000 lb. is again resolved into a direct
pull along the axis of the leg amounting to 42,000 lb., and a
shear of 13,500 lb. The area " BB " being 64 in.^, therefore
the direct tensHe stress at " BB "=42,000-1-64 = 657 Ib./in.^
The shear load of 13,500 lb. at " BB " being opposed by an
equal but opposite shear load at " AA " there must be a
couple of magnitude 13,500 x 24 in. lb. to keep the part of the
leg lying between " AA " and " BB " in equilibrium. Assum-
ing that this couple is composed of two equal parts operating
at "AA"and"BB,"
Bending moment at "BB" = 13,500x24^2=162,000in. lb.
Section modulus ,, ='^170in. ^
Bending stress „ =162,000-:-170 = 953lb.per sq.in.
Total tensile stress . .= 953 -|- 657 = 1610 „ „ „
The stress in the studs depends upon the degree to which the
nuts are tightened, and if of sufficient area the stress is prob-
ably not affected by the applied load. It therefore only remains
to see if the loads induced by tightening them up to their
nominal working stress are sufficient to prevent the joint
opening.
Referring to the table on page 130, the nominal load of five
If in. studs = 75,000 lb. Subtracting 44,000 lb. there remain
31,000 lb. to resist tilting about the outside edge of the foot.
The distance from to the centre of mean position of the studs
is BJin., and the corresponding moment is therefore 31,000 x 6-5
=202,000 in. lb., which is greater than the moment to be
resisted.
Design of Crank-cases . — ^A simple form of crank-case is shewn
on Fig. 93, stroke to bore ratio 1-25. The crank-case is machined
on each side and in position is bolted to its neighbours, an
arrangement which, though ^nusual, has its advantages, both
in the factory and outside, as the small sections are easier to
handle than a crank-case in one piece. Provision is made for
staybolts, and the thicknesses of metal shewn are about the
minimum to give satisfactory results with this design. It must
not be thought that these proportional thicknesses are capable
of substantial reduction with large sizes of engine without
modification to the distribution of metal. On the other hand,
if the interior of the case is built up in girder or box formation,
FRAMEWORK
137
or generally reinforced by internal ribbing, as shewn dotted in
Fig. 93, the general thickness may be reduced by about 25%,
and advantage is taken of this fact in designing large engines,
the castings of which would be of undesirable thickness if the
simpler forms were adopted. In the event of staybolts not
being used, it is desirable to check the stresses in the legs in the
manner described for an " A " frame. Judging by successful
138
DIESEL ENGINE DESIGN
designs, the use or omission of the staybolts has little influence
on the strength of crank-case required, and this is readily
explained by the fact that whereas the staybolts relieve the
crank-case of tensile stresses, the tightening of the former
throws a heavy buckling load on the crank-case, perhaps
double the tensile load due to the working pressure in the
cylinders. These considerations do not apply, however, to
those designs in which the staybolts are extended upwards to
the cylinder cover. In these cases the crank-case has only the
guide pressure to contend with. On the other hand, the use of
Fig. 94
FRAMEWORK
139
staybolts passing through lugs on the cylinders enables the
latter to be pitched closer together than would be easily possible
otherwise, and in any case they add strength and rigidity at
very little expense and increase in weight.
The above notes on the strength of crank-cases, as well as
the figures for the thickness of metal, apply equally to cross-
head as to trunk engines. The additional height of the former
has little if any influence on the matter, as the guide reaction
acts at the same height, above the centre line of the crank-shaft,
assuming the same length of connecting rod in each case.
Some alternative forms of construction are shewn in Fig. 94.
In the case of very small engines the use of the minimum
thickness of metal allowable on considerations of strength and
rigidity only, would give rise to trouble in the average foundry
and additional thickness is usually given. In the following
table foundry considerations are neglected, as these must be
dealt with by the designer in each individual case, in accordance
with his judgment of the capabilities of the foundry in question,
and in this matter the foundry manager will be able to give
assistance.
General thickness of
General thickness of
Diameter of
Whit-worth
Bore of
Crank-case metal.
Crank-case metal.
Cylinder,
in.
m.
Plain type,.
m.
Box or girder formation.
Staybolts,
in.
Fig. 93.
Fig. 94. 1
10
i"
tV"
ir
12
r
r
If"
15
i"
x%"
H"
18
i"
i"
2i"
21
ItV"
I"
2|"
24
ifV"
I"
H"
27
If"
ItV"
3|"
30
H"
ifV"
3|"
The figures for the diameters of the staybolts are based on the
assumption that they carry the whole pressure load. In cases
where the cylinders are secured to the crank-case by a studded
flange the staybolts if fitted at all may be made considerably
lighter, according to judgment or the results of experiment.
Other points to be considered in designing a crank-case are : —
140 DIESEL ENGINE DESIGN
(1) The provision of oil-tight access doors of ample size for
overhauling the bottom ends.
(2) End casings provided with oil flingers, stuffing boxes, or
other means of preventing the escape of oil.
(3) Facings, and other necessary accommodation for valve
gear.
(4) Bosses to carry lubrication oil connections to the main
bearings.
(5) Facings for platform brackets.
(6) A vent pipe or valve of large area, to relieve pressure in
the event of an explosion in the crank-case without loss
of lubricating oil during normal working.
(7) Steady pins to each section of the case, to fix correct
location.
Machining the Framework generally. — In designing all parts
of an engine the designer will keep in mind the capabilities and
limitations of the manufacturing plant and the operatives.
This is especially necessary in the case of the framework, on
account of the relatively large size of the parts. Where the
most modern type of face milling plant is available the element
of size offers no difficulties, and bedplates of 60 feet in length
may be faced in one operation. Where planing must be resorted
to the capacity of the machines must be studied in the early
stages of the design. Machined faces should be arranged in as
few different planes as possible, and ribs or flanges projecting
beyond those planes are to be avoided as much for convenience
in machining as for the sake of appearances. The simpler forms
of girder or box-girder construction are to be preferred to those
designs in which alternate perforation by lightening holes and
reinforcement by ribbing mutually defeat each other's object.
The lightest, strongest and cheapest forms are to be attained
with a minimum of holes and ribs when cast iron is used. Large
steel castings, however, are preferably lightened out almost to
the extent of lattice-work, in^brder to facilitate rapid stripping
of the cores after solidification and to minimise initial stresses.
Literature. — The different types of framework used in Marine
Diesel Engine construction, and the forces acting on them,
are discussed in the following paper : — Richardson, J., " The
Development of High Power Marine Diesel Engines," Junior
I.E., April 20th, 1914.
CHAPTER VIII
CYLINDERS AND COVERS
General Types. — ^The great majority of Diesel Engines are
provided with cyliader liner, jacket and cover as separate
pieces, as in Fig. 95, which refers to a four cycle trunk engine.
Different arrangements have, however, been used successfully,
and deserve mention. With small engines, simplification is
achieved by casting the jacket and liner in one piece, as in
Fig. 96. Remembering that the bulk of the jacket wall remains
Fig. 95.
Fig. 96.
141
142
DIESEL ENGINE DESIGN
stone cold, it will be appreciated that this construction involves
increased tensile stresses on the jacket, due to the tendency of
the liner to expand, and jackets of this kind have been known
to crack circumferentially. In cases where staybolts have been
fitted to carry the tensile stresses from the cover downwards
little damage has resulted. On the other hand, when the jacket
has been relied on for this function, rupture during working
may easily occur, and has sometimes resulted
in the cylinder being projected towards the
roof. These considerations would appear to
indicate that the use of this construction,
"without staybolts or other safeguards, is not
lightly to be attempted without serious con-
sideration of the capabilities of the foundry.
In Fig. 97 is shewn a construction in which
the cover is incorporated with the cylinder
casting in motor-car style. In this case the
tensile stresses are mainly carried by the liner,
and the jacket is made relatively thin and
flexible. This design, though probably safer
than that of Fig. 96, also makes some demand
on the skill and care of the foundry people.
In this connection it is worth while bearing in
mind that many failures might possibly have
been avoided if it had been realised that
certain continental designs, in which lightness
has been, a primary consideration, were only
practicable if the greatest care were exercised
in the selection of material and in making the
castings. There are other types of cylinder in
successful use, notably .nose in which the liner
and cover are cast together apart from the
jacket, but the remainder of this chapter will
be devoted tcjthe consideration of the details
of the more common construction, in which the liner jacket
and cover are separate pieces. Unless the contrary is stated,
cast iron is understood to be the material in each case.
Liners. — Special cast iron is used for liners, but there is little
unanimity of opinion as to the most desirable properties be-
yond the obvious requirements of soundness and homogeneity.
The greatest difficulty to be overcome is abrasion by the piston
rings. At present it seems open to question whether the
Fig. 97.
CYLINDERS AND COVERS
143
problem is most influenced by the material of the liner or the
piston rings themselves. It is very seldom that liners crack*
either in four or two stroke engines, and this immunity is
doubtless due to the rapid conduction of heat from the breech
end to those parts which only come in contact with gases at
relatively low temperatures, from which it would appear that
the best way to cool a relatively inaccessible spot is to connect
03 to 'OS 8
'Atofs^
Fig. 99.
the latter to a large, well-cooled area, as near to it as possible.
This principle will be referred to again in connection with
exhaust valve casings.
Typical liners for four stroke engines of the trunk and cross-
head types respectively are shewn in Figs. 98 and 99. With
the latter the piston is only of sufficient length to carry the
rings, and the length of the liner is determined by the position
of the bottom ring at the bottom of the stroke . With the trunk
engine the liner must be long enough to embrace a sufficient
* Except at a piston seizure.
144 DIESEL ENGINE DESIGN
length (about equal to the bore) of the parallel part of the
piston when at the bottom of its stroke, in order to avoid a
piston knock at the bottom dead centre. It is therefore
necessary to determine the clearance volume and complete the
design of the piston before the length of the liner can be fixed
finally.
The bore is usually parallel with four stroke engines and
slightly barrelled in way of the ports in two stroke engines to
allow for the restraints which are inevitably placed at that
position against free expansion of the liner. Probably the best
bore is produced by finishing with a reamer in a vertical
machine. Grinding is frequently adopted, but there is a
question if this prpcess does not to some extent destroy those
properties of cast iron which facilitate good lubrication. The
outside surface is frequently left unmachined in competitive
work, and there is probably no serious objection to this practice
for four cycle work. For two cycle engines it seems reasonable
to take advantage of the increased heat conductivity obtain-
able by removing the skin.
Strength of Liners. — ^The upper end of the liner is subject to
a working pressure of about 500 lb. per sq. in., and the thick-
ness at this part measured under the heavy top flange may be
found by the following formula, which represents average
practice for substantial engines : —
Thickness =0-08 bore -|- J"
The working stress being about 3000 lb. per sq. in. in the
case of a 30 in. cylinder, and less in smaller sizes ; explosions at
starting, etc., may nearly double this stress occasionally.
Unfortunately the available information on the effects of
repeated stress is not sufficiently complete at present to enable
one to say definitely whether or not these excesses of stress
have any influence on ultimate failure by fatigue, but the
writer is inclined to believq|(on the strength of such evidence
as has come before his notice) that the elimination of these
occasional excess pressures would not enable any substantial
reduction of thickness to be effected with the same margin of
safety.
On account of the diminution of pressure on expansion the
liner may be tapered to a thickness of about 0-04 bore at the
open end.
The breech end of the liner requires to be reinforced by a
CYLINDERS AND COVERS 145
heavy flange, to avoid distortion due to the pressure of the
cover on the spigot joint. Proportions are given in Fig. 98.
Points of Detail. — The difficulty of accommodating the
valves in the limited space available in the cover of a four
stroke engine usually renders it necessary to make recesses in
the top of the liner to clear the air and exhaust valve heads
(see Fig. 98). Four or more tapped holes are provided in a
circumferential line round the liner to accommodate the
lubricating fittings, these being drUled when the liner is in
position in the jacket. The holes are located at about the level
of the second piston ring (counting from the top), when the
piston is at the bottom dead centre. The fittings themselves
will be described later. The water-joint between liner and
jacket at the lower end may be made by one or more rubber
rings. The joint between cover and liner may be of copper or
asbestos compositions.
Two stroke liners are complicated by exhaust, and some-
times air ports (see Fig. 102). In the earlier designs the bars
between the latter were always provided with water passages,
which introduced difficulties in manufacture, and the value of
which seems doubtful, and these are now frequently omitted.
The fitting surfaces at this point are preferably ground, to
minimise chance of leakage, as the high temperature prohibits
the use of rubber packing rings.
Cylinder Jackets. — A simple and effective form of jacket for
a four cycle engine is shewn in Fig. 95, and in this example the
jacket takes the pressure pull without the assistance of stay-
bolts. The chief points to be observed are : —
(1) A heavy flange at the top to carry the liner and to enable
the tensile forces concentrated at the studs to dis-
tribute themselves uniformly round the jacket without
producing high local stresses.
(2) A nearly plain cylindrical barrel, as nearly as possible in
line with the pitch circle of the cover studs and provided
with sludge doors, bosses for lubricating fittings, and a
bracket for supporting the cam-shaft bearings.
(3) A circular flange at the bottom for securing to the crank-
case.
The remarks re tensile forces under heading (1) apply here
also, but to a less degree, as the studs are pitched closer to-
gether than would be feasible on the cover. On these consider-
146
DIESEL ENGINE DESIGN
ations the thickness of the jacket for eoual strength should
taper gently towards the middle, and the form shewn in the
figure is the practical compromise. Some of these points will
be considered in greater detail.
Top Flange of Cylinder Jacket. — In small engines this may
be soUd, but with larger sizes, say from 15 in. bore and upwards,
difficulty is sometimes experienced in obtaining sound metal
at this point, and coring of the flange between the studs is
resorted to in order to accelerate cooling in the mould.
Different constructions are shewn in Fig. 100. Schemes A and
B have the additional advantage of increasing the cooling
surface. Where four cycle engines are concerned the im-
portance of this consideration is probably negligible. Scheme
B requires a water outlet connection between each stud if air
pockets are to be avoided. On the other hand, the expense of
coring is less than with scheme C.
Barrel of Cylinder Jacket. — ^This is sometimes conical,
instead of cylindrical, and in*his case it is reasonable to provide
a vertical internal rib under each stud to discount the addi-
tional stress involved. Consideration of manufacturing costs,
and of the good appearance of the engine, rule out of court any
form of external ribbing. The brackets supporting the valve
gear take many forms in different designs. That shewn in
Fig. 95 is the modern form, and considered in conjunction with
the gear it supports appears to combine most advantages,
including that of elegance.
CYLINDERS AND COVERS
147
Bottom Flange of Jacket. — If four staybolts are provided for
each cylinder, these may conveniently be used to secure the
latter to the crank-case. The concentration of the tensile load
at four points necessitates a heavy flange, preferably of box
form, as shewn in Fig. 101. I .
The corners of this flange 8-/^5turfs, ..—"0^
being each subject to a
load of one-quarter of the
maximum working pres-
sure load, deserve atten-
tion in the form of a
calculation of the bending
stress involved. A plain,
square shape would appear
to be preferable to some of
the more elaborate shapes
which have occasionally
been used, the flat sides
lending themselves well to
the provision of facings
for various purposes.
Frequently the flange
is spigoted into the top
of the crank-case, but as
this involves an unneces-
sary machining operation
on the latter and makes
cylinder alignment more
difficult, the better prac-
tice is to core the aperture
in the crank -case suffi-
ciently large to allow for
adjustment of the position
of the cylinder and to locate the latter by means of two
steady pins.
Strength of Four Stroke Cylinder Jackets. — ^The consider-
atidns of strength which enter into the design of a cyhnder
jacket are illustrated by the following check calculations
relating to the cylinder shewn in Fig. 101.
Bursting stress in liner
500 (lb. per sq. in.) X 6-75 „„„„„
= ^ ■, -.2^ — =3000 lb. per sq. m.
1-125 ^ ^
Fig. 101.
148 DIESEL ENGINE DESIGN
■vr • 1 11 • u ^ J 500x0-784x142 „...,,
JNominal pull in each cover stua=^ =9650 lb.
o
Permissible nominal load for 1| in. stud, according to table
on page 130, 9300 lb.
Maximum working pull in jacket=0-784x 13-52x500 =
71,5001b.
71 500
Tensile stress in jacket= ' — pr-;7^=1670 lb. per sq. in.
Owing to the peculiar shape of the bottom flange the calcu-
lation of its strength presents a difficulty which is easily evaded
by substituting for the actual section a simpler one of obviously
inferior strength.
Nominal load at each corner, 71,500 lb.-^-4 = ~18,000 lb.
Moment from centre of bolt to jacket wall=18,000 x 9 in. lb.
Modulus of hypothetical section : —
^^/10-5X53_9-5X3-53N ^,^^3^.^.^3
o. 18,000X9 '. ;- .nAii.
.•. Stress< — ^--— — I.e. <5,400 lb. per sq. m.
In view of the unfavourable assumptions this is probably
not excessive for first-class cast iron.
Jackets for Two Stroke Engines. — The necessity for provid-
ing exhaust passages or belts, and in some cases passages for
scavenge air as well, introduces considerable complication into
the design, renders the stresses in certain parts more or less in-
determinate, and makes greater demand on the skiU of the
manufacturing departments, in comparison with that required
by four cycle construction.
Referring to Fig. 102, it will be seen that the exhaust belt
interrupts the vertical line of the jacket wall, and if the latter
has to carry the main tensile stresses internal ribbing becomes
a necessity. The arrangement shewn is perhaps as good as any,
but the attachment of ribs to*he exhaust belt has a restraining
influence on the temperature expansion of the latter which
can only result in mutual stresses. It appears, however, that
these are not very serious, as cylinders which have failed in
other respects have remained intact at this point. Fig. 103
shews a construction in which a good attempt is made to
secure continuity of the vertical wall of the jacket. Either
of these systems is probably satisfactory for cylinders of
medium size. Large cylinders, however (and this applies to
Fig. 102.
Fie. 103.
Fig. 104.
150
DIESEL ENGINE DESIGN
other parts as well), are known to be subject to greater temper-
ature differences than smaller ones (though not to the extent
sometimes suggested,), and the leading designers have had
recourse to other expedients when faced with the problem of
constructing cylinders of large size.
In Fig. 104 the jacket wall may be described as similar to a
honey -pot in shape and of abnormal thickness, to allow for the
bending stresses caused by the curvature of the walls and the
fact that the tensUe supporting forces are localised at two feet.
Cylinder Lubrfcating
Oil Rinii Main. bl, , . .. „
-" T^Lubncating Oil Inlet
Fig. 105 (See page 152).
The exhaust belt is of relatively thin metal, with comparatively
small support from the walls. It wUl be evident that the
strength of the jacket is very slightly influenced by the exhaust
belt, and that the latter is free of all but temperature stresses.
This construction, therefore, Sttains a good approximation to
the correct allocation of the respective duties of jacket and
exhaust belt.
As disadvantages, may be cited abnormal weight of cylinder
and the difficulty of casting a cylinder involving widely
different thicknesses of metal.
Another and perhaps better way out of the difficulty is to
connect the cylinder cover to the bedplate by means of stay-
CYLINDERS AND COVERS
151
bolts, thus relieving the jacket of all stresses except those
induced by temperature differences. The jacket in this case
virtually hangs from the cylinder cover, and only requires to
be attached thereto by studs proportioned to a load based
on the cyhnder pressure and the annular area lying be-
tween the cylinder bore and the spigot at which the cover
joint is made. The upper flange is preferably made fairly
substantial, but other thicknesses may be made a practical
minimum.
Cylinder Lubricating Pump.
Sti"ofee
Hole toprevent
AirLock ^
Driven by Eccentric
or LinkXvork
.t— i Plunders arranged
i 7n a Ore le
Suction fei
by drip Lubricator
^Delireries *
Fig. 105 (See page 152)
Cylinder Lubrication. — ^The problem of cylinder lubrication
in Diesel Engines consists in effecting uniform distribution of
minute quantities of oil. The quantity of oil admitted must
be the minimum necessary to effect satisfactory lubrication, as
the oil " cracks " in service, leaving a gummy deposit, which in
course of time causes the piston rings to stick. Under favour-
able conditions this may be several months, even a year.
Every drop of superfluous oil reduces this period, hence the
importance of uniform distribution so that every part may
have sufficient, but none a superfluity. These conditions are
152
DIESEL ENGINE DESIGN
best secured, by a separate controllable feed to each of about
six or eight points round the circumference of the cyhnder. A
typical lubricating fitting is shewn in Fig. 105 (pages 150-152),
and the point to be observed is that the fitting must adapt itself
to slight relative movement between the liner and jacket. The
small hole at the end which leads to the surface of the liner
reduces to a minimum the chances of the fitting becoming
choked with carbon. With forced lubricated engines, in which
the cylinder is not isolated from the crankpit, it frequently
happens that more than sufficient oil reaches the cyhnder, apart
from any arrangements made for the purpose. In this case the
problem may be to devise scraper rings, vent holes, or other
devices, to remove the superfluous oil.
Liner
^Jacket
Cylinder Lubricating
Fitting.
Fig. 105 (See also pages 150, 151).
Cylinder Covers. — Owing to a considerable number of
failures in service and difficulties experienced in manufacture,
cylinder covers for both four and two cycle Diesel Engines
have come to be regarded as difficult pieces of design, and it
may perhaps be instructive to review the subject in a more or
less historical manner.
The earlier type of four c^le cover is shewn in Fig. 106,
from which it will be seen that the internal coring is compli-
cated and that a few core-holes of small diameter only are
provided to vent the core in the mould. In spite of these dis-
advantages, such covers have given good service when made
by the most skilful of continental manufacturers. Dismissing
for the moment the question of manufacturing costs, these
covers have the following shortcomings : —
CYLINDERS AND COVERS
153
Fio. 106.
154
DIESEL ENGINE DESIGN
( 1) The thin walls of uncooled, metal between the recesses for
fuel and, exhaust valves are liable to crack on over-
loaded engines.
(2) The hot exhaust passage is too rigid to permit of much
expansion and leads to cracking of the bottom plate.
(3) The small core-holes give poor access to the interior for
purposes of cleaning away accumulated scale.
Assuming first-class foundry work, the two latter considera-
tions are perhaps the most important. Modern development is
on the following lines : —
The provision of large doors, which serve the double
purpose of providing good access for cleaning and
(1)
affording better support and venting for the core when
casting.
(2) Elimination of all internal ribs, as experience seems to
shew that the tubular walls provided to accommodate
the valve casings provide all requisite support between
the top and bottom plates.
(3) Using brass or steel tubes expanded into the recesses for
the fuel valve and holding-down studs.
(4) The use of square instead of conical seats for the valve
casings.
Fig. 107.
A cover designed on these lines is shewn in Fig. 107. Some
makers have simplified the question of casting at the expense
of introducing extra machining and fitting operations by
making the top plate a separate piece (see Fig. 108.)
Another innovation which is becoming increasingly common
is to place the fuel valve off centre. This arrangement enables
CYLINDERS AND COVERS
155
the cooling space around, the fuel valve to be increased, but too
great a displacement of the fuel valve from the centre position
necessitates a special shape of combustion space.
Points of Detail. — Owing to the large recesses for the valve
cages, a four cycle cover is relatively weak, considering the
amount of metal in it, and on this account all stud holes should
be well bossed under and all inspection openings well reinforced
by compensating rings like a boiler.
The under face of the cover is machined all over, but on the
top face machining is sometimes restricted to those parts which
are occupied by valves, etc. This enables the corners to be
given a liberal radius, which in addition to improving the
Fig. 108.
appearance facilitates moulding (see Fig. 107). From all
considerations, all internal angles should be well radiused.
Water is led to the cover by one of two methods : —
(1) By one or more tubular fittings screwed into the top of
cylinder jacket and passing through holes in the under
face of the cover (see Fig. 109). With the type ot
jacket shewn in Fig. 100b it is desirable to fit one such
fitting between each pair of cover studs.
(2) By means of an opening in the side of the cover (see
Fig. 110).
Whatever means be adopted, it is advantageous to fit
internal pipes or baffles to encourage fiow towards the
fuel valve, as accumulation of deposit at this point is to be
avoided at all cost. It is usual to arrange the outlet above the
exhaust branch, as stagnation at this point is also undesirable.
156
DIESEL ENGINE DESIGN
Proportions of Cylinder Covers. — ^The depth of a four stroke
cover generally works out to about 0-7 of the cylinder bore,
the limiting factors being the size of the exhaust passage and
the water space around it. The former should be at least
equal in area to the exhaust valve at full lift. The passage
starts by being rectangular in shape at the valve end, and
gradually becomes circular at the outlet where the diameter is
about 0-31 of the cylinder bore. The same applies to the air
inlet passage. The thicknesses of metal vary considerably in
different designs, and the proportions shewn on the sketches
represent average practice. The tendency seems to be to
Fig. 109.
Fig. 110.
rely for safety on a good thickness of metal on the bottom
plate.
Strength of Four Stroke Cylinder Covers. — The system of
loads acting on the cover comprises the tightening stresses of
the studs, the reaction at the spigot joint and the gas pressure
on the lower plate. The effect of such a system is to produce
tensile stress in the top plate and compression on the bottom.
Considering the relative weakness of cast iron in tension and
the fact that cracks in the td|) plate are of very rare occurrence,
it would appear that covers proportioned in accordance with
average practice have a good margin of safety so far as pressure
stresses are concerned. In view of a few isolated failures, or
rather as a matter of principle, the strength should be subject
to calculation. Owing to the uncertainty as to actual condi-
tions the method of calculation detailed below must be con-
sidered comparative rather than absolute.
CYLINDERS AND COVERS
157
The assumptions underlying the method are as follows : —
(1) That the severest conditions of stress are due to a cylinder
pressure of 1000 lb. per sq. in., due to pre-ignition,
careless starting or otherwise, and that this pressure
causes the cover to lift to such an extent that the
reaction at the joint is eliminated.
(2) That the stress is uniform across a diametrical section in
the case of a cover of constant depth and proportional
to the distance from the neutral axis of the section in
the case of a cover of varying depth. This is not correct,
but probably involves approximately equal percentage
of error in different cases.
Fig. 111.
Example : Referring to Pig. Ill, shewing the weakest
section passing through the recesses for the air and exhaust
valves, the section modulus is 114 in. ^
Considering the forces to the right or left of this section,
we have : —
(1) A downward force at the stud circle equivalent to a
pressure of 1000 lb. per sq. in. over half the circular area,
extending to the joint spigot, viz.: -784 x 13-252 x
1000-^2 = 69,000 lb. This may be considered to act at
the centre of gravity of the pitch semicircle, that is at
a distance of 9-5x2^7r=6-02 in. from the section
under consideration.
158
DIESEL ENGINE DESIGN
( 2) An equal and, opposite force on the under side of the cover
acting at the centre of gravity of the semicircular area
extending to the joint spigot, i.e. at a distance of
6-625x4
3n
=2-8 in. from the centre.
The stress is therefore : —
69,000 (6-02-2-8)
114
' = 1,950 lb. per sq. in.
Putting the above into the form of a rule :-
f =
1000 R^g (Bz-f Ri)
Where 1000= Assumed maximum pressiu-e.
f =Stress in lb. per sq. in.
E,2=Ba'dius of stud pitch circle in in.
Ri=Inside radius of joint ring in in.
z=Section modulus in in. *
Two Stroke Cylinder Covers. — ^Where port scavenge is
adopted the cover has only to accommodate the following
fittings : —
(1) Fuel valve. (3) Relief valve (if fitted).
(2) Starting valve. (4) Indicator tube fitting.
Fig. 112.
CYLINDERS AND COVERS
159
As all the above are relatively small the construction of the
cover is much simpler than that for a four stroke engine, and
failures are rare. In addition, the absence of air and exhaust
valves enables the lower plate to be dished upwards (see
Eig. 112), an arrangement which gives greater freedom of
expansion and a shape of combustion space which is probably
favourable to combustion. Eig. 112 shews a cover of this type
arranged for four staybolts.
When valves in the cover are employed for scavenging
" ^A^W^^^^^^
Fig. 113.
purposes the construction depends on the number of valves.
If two scavenge valves are used the cover may be of the four
cycle type and interchangeable with those of four cycle engines
of the same bore. When fitted to a two cycle engine the
average temperature of the lower plate will be higher ; but on
the other hand the temperature distribution will be more
uniform than that of a similar cover fitted to a four stroke
engine where one of the valves is used to conduct exhaust gases.
The inequality of temperature over the bottom plate of a four
stroke cover is strikingly illustrated by the observed fact that
in the event of failure the first crack almost invariably runs
160
DIESEL ENGINE DESIGN
between the pockets for the exhaust and fuel valves. A crack
between the air valve and, the fuel valve sometimes develops
at a later stage.
Fig. 113 shews a cover designed to accommodate four
scavenge valves. It will be noticed that the interior is divided
by a horizontal diaphragm separating the air space and the
water-jacket. It appears that this diaphragm and the tubular
connections to the bottom plate impose too great restrictions
on the expansion of the latter, and fractures have been fre-
quent (with both cast iron and cast steel), so that this type
of cover, as hitherto designed, must be considered a failure.
The writer understands that a modification of this design
(patented by Mr. P. H. Smith), shewn in Fig. 114, has proved
satisfactory, and at present no failures have been reported.
Apparently the additional depth of the water-jacket, and
correspondingly increased freedom of expansion, minimise
temperature stresses, and the support afforded by the ex-
ternal shell keeps the bending stresses to a moderate figure.
FiQ. 114.
Literature. — "Diesel Engine Cylinder Dimensions." — Article
in Engineering, September 5th, 1913,
Richardson, J. — Paper on Marine Diesel Engines, loc. cit.,
p. 64.
Hurst, J. E. — " Cast Iron with Special Reference to
Engine Cylinders," Manchester Assoc. E., December 9th, 1916.
CHAPTEB IX
RUNNING GEAR
Trunk Pistons. — ^These are so well known in connertion with
petrol motors, gas engines and, the like that a general descrip-
tion is unnecessary, and we may at once proceed to the con-
sideration of their special requirements in Diesel Engine con-
struction. The difficulties involved in combining the piston
proper with the crosshead arise chiefly from the heat which
reaches the gudgeon pin bearing by conduction and radiation,
and the high pressures dealt with in Diesel Engines (as com-
pared with gas engines) necessitate a high specific pressure at
this bearing owing to the limited space available. The most
serious troubles to be anticipated are piston seizures, which
like all other heat troubles are more pronounced in large than
in small engines. For these reasons trunk pistons are not
generally used for cylinders exceeding about 22 inches in
diameter. For marine engines the inaccessibility of the
gudgeon pin bearing is usually considered an insuperable
objection for all but the very smallest cylinders. It is possible
that prejudice has a little to do with this view, and it is interest-
ing to note that Semi-Diesel Marine Engines of fairly large size
(500 H.P. for example) apparently give good service with
trunk pistons. It is evident that a really efficient system of
water cooling would abolish most of the difficulties mentioned.
Unfortunately such systems as hitherto fitted to trunk engines
have not all been uniformly successful. On the other hand a
great measure of success has been achieved in some cases and
the ultimate solutions (if not already achieved) are probably
near at hand. For crosshead engines water cooled pistons on
two or three different systems have already emerged success-
fully from the experimental stage.
Material. — The use of cast iron for pistons is almost universal
on account of its good wearing properties and its cheapness.
Owing to the low guide pressure the quality of the metal is
M 161
162
DIESEL ENGINE DESIGN
probably of minor importance so far as wear is concerned,, as
the latter is in any case hardly measurable. On the other hand,
so far as that portion of the piston is concerned which is in
contact with the working fluid (viz., the piston crown), the
quality of the metal is of great importance in determining the
liability or otherwise of the crown to crack under the influence
of heat. According to Mr. P. H. Smith, the irons whicli give
the best results are those of the coarsest possible grain.
Experience in other directions seenis to indicate that a large
carbon content and low percentage of phosphorus are
favourable.
At the first glance it might appear strange in view of the
^
^
^ 1^
^
1
D.
^
.?
£ y F.
X
Fis. 115.
fact that heating produces compression at the point of maxi-
mum temperature, that the cracks start at the centre of the
crown on the side in contact with the gases, but -the facts are
easily explained by some such hypothesis as the following : —
The local heating causes local compressive stress of high
intensity which in course of time causes the particles to re-
arrange themselves in such a iaanner that this stress is reduced.
On cooling the contraction of the surrounding metal induces
tensile stresses in the centre equal in amount to the extent by
which the original compressive stress has been reduced. Con-
siderable support is afforded to this theory by the fact that the
cracks develop in the course of time into fissures, shewing that
the material has contracted circumferentiaUy. Also the
plastic deformation of cast iron at a red heat is often observed
RUNNING GEAR
163
in such familiar articles as kitchen ranges and the like, in which
no special provision is made for expansion or growth.
Shape of Piston Crown. — ^Fig. 115 shews some alternative
shapes.
Of these types C is the best for combustion, though types
A and B are little inferior in this respect. With existing
methods of fuel injection, type E is quite inadmissible on the
score of combustion unless the cylinder cover is concave
downwards, and even then its efficiency is very doubtful. On
the other hand, type F, with the shape of combustion space
indicated appears to give good results, due doubtless to the
manner in which the charge of air is concentrated. A little
reflection shews that the curved shape of type B gives rise to
greater compressive stress at the centre on the combustion
side than type A on account of the bending action which
arises when a state of temperature stress comes into existence.
Type D is very liable to failure unless a spreading flame plate
is used. From this it would appear that of aU the types
illustrated B and D are the most liable to failure. In practice
types B and C are the most usual, and their inherent weakness
in the larger sizes (small pistons rarely crack) is guarded against
(1) Careful selection of material.
(2) Providing a considerable thickness of metal.
(3) Water cooling.
(4) Providing loose crowns, or central cores.
The provision of a great thickness of metal assists matters
in two ways : —
(a) By reducing the bending
stresses due to both tem-
perature (probably the
most vital) and pressure.
By giving additional area
for heat flow to the walls
of the liner.
(b)
Fig. 116,
Fig. 116 is an attempt to give a
diagrammatic representation of the
heat flow by increasing the inten-
sity of shading towards the parts
having the higher temperatures.
The usual form of water cooled
Fig. 117.
164
DIESEL ENGINE DESIGN
trunk pistons is shewn in Fig. 117, and the means adopted
to convey the water to and from the water space wiU be
discussed later. In the earlier types the space was intended
to be full of water, but in more recent designs the tendency
is to rely on a smaU flow of water, part of which is evaporated
and absorbs a relatively
large quantity of heat
in latent form.
Fig. 118 shews a loose
piston top secured by
four studs, the holes for
which have sufficient
clearance to allow for
the expansion of the
former. Fracture of
such a loose top due to temperature efiects is improbable,
and if it occurs the cost of renewal is trivial.
Fig. 119 shews a similar design patented by Mr. P. H.
Smith and applied by him to pistons in which cracks had
Fia. 118.
Fib. 119.
already appeared.
Proportions of Trunk Piston
Bodies. — The usual proportions
found in practice are discussed
below with reference to Fig. 120,
which is purely diagrammatic.
The thickness (C) of the crown
(with uncooled pistons) increases
rapidly as the bore is increased,
for reasons which have been noticed
above, and the following figures are
a guide to good practice : —
Bore of cylinder Thickness C
10" .
. li"
12" .
. If"
14" .
. 21"
16" .
. 3"
18" .
. 4"
20" .
. 5"
Fio. 120.
The distance from the top of the
piston to the first ring may con-
veniently be made equal to C. With
RUNNING GEAR 165
the above values of C this prevents the first ring being placed
in too high a position where its proximity to the source of heat
would cause it to become stuck with carbonised lubricating
oil in a short time. The rings themselves may be of square
section with R=0-025 to 0-033 B. The gap between the
ends of the ring when free may be about 2-50 times R^.
The number of rings fitted varies from about five in
small to eight in large cylinders, and the space between
consecutive rings is not as a rule less than R. The construc-
tion of piston rings is rapidly be-
coming a specialised branch of in-
dustry and details will not be given
here. A method of designing and
manufacturing piston rings is de- ' j.^^ ^21
scribed by Guldner in considerable
detail ("Design and Construction of Internal Combustion
Engines "). It is important to prevent the joints of the rings
working into line during service, and a method of fixing
them is shewn in Fig. 121.
Dimension " A " varies greatly in different cases. The
maximum value found in practice, viz. A=2B gives a piston
of ideal running properties, but is seldom fitted nowadays
owing to considerations of first cost. A=1-4B to 1-6B repre-
sents good average practice. Still smaller values of A are
sometimes used, but are not to be recommended. The pro-
vision of a long piston skirt is advantageous from the following
points of view : —
(1) Rendering possible a low and therefore comparatively
cool location for the gudgeon pin bearing.
(2) It minimises piston knocks.
(3) It facilitates the flow of heat away from the crown by
providing a large surface in contact with the cylinder
walls.
The upper part of the piston body is turned taper to allow
for expansion, and the approximate allowances to be made on
the diameter are given in Fig. 122, in which the taper is greatly
exaggerated. The diameters of the piston ring grooves are
figured by allowing all the grooves to have the same depth
below the tapered surface. The clearance behind the grooves
1 R denotes the thickness of the ring measured radially. The bars K between
the rings may be equal to or slightly greater than R.
166
DIESEL ENGINE DESIGN
should be a practicable minimum. The gudgeon pin is usually
located either at the centre or slightly above the centre of the
parallel part of the body, and allowance is made for possible
local distortion due to driving in the gudgeon pin or to ex-
omtoo-ozo
Fig. 123.
Fig. 122.
pansion of the latter by relieving the piston surface to the
extent of about 20-thousandths of an inch, in the manner
indicated in Fig. 123.
Approximate figures for other main proportions are given
below : —
E=2G.
G=0-4B.
H=0-5B.
C (see Table above).
D=0-07B.
E=0-033B.
Gudgeon Pins. — Alternative forms of gudgeon pins are
shewn in Fig. 124, type A being most generally used. The pin
itself is of special steel, case-hardened and ground if working
in a bronze bearing. If the bearing is white-metaUed the case-
hardening is unnecessary. The pin is a driving fit in the body
and is secured in the mann# indicated in the illustrations.
Lubrication of the pin may be effected in various ways : — -
(1) Forced lubrication by means of a pipe or drilled hole
leading from the big end of the rod.
(2) By means of a groove or pocket on the surface of the
piston communicating with the surface of the gudgeon
pin and fed by means of a fitting similar to that shewn
in Fig. 105 (see Fig. 125).
RUNNING GEAR
167
Assuming the proportions given in the previous article, the
maximum bearing pressure works out to about : —
0-785x500 , xor^nmi,
=about 2000 lb. per sq. m.
0-4X0-5
Fig. 124. Type A.
and it is hardly surprising that good bronze appears to be
preferable to white metal for this bearing.
Tig. 124. Type B.
168
DIESEL ENGINE DESIGN
Miscellaneous Points of Detail. — ^Where forced lubrication
is employed it is important to prevent oil from splashing on
to the hot piston crown and some arrangement of baffles is
very desirable. This may take the form of a plate across the
mouth of the piston body or a light guard over the cranks. A
piston ring at the lower end of the piston is useful in removing
superfluous oil from the liner and in effecting a good distribu-
tion of the film.
It is generally necessary to cast two shallow recesses in the
/l i-X.
H
Fig. 126.
Fig. 125.
piston crown to clear the air and exhaust valve heads at the
top dead centre, and the positions of these being remote from
the point of greatest tempera||ire'are usually chosen for two
tapped holes to receive lifting bolts. Holes should not be
drilled in the centre of the crown, and if a turning centre is
necessary a special boss should be cast for this purpose and
turned off afterwards (see Fig. 126).
So far nothing has been said about pressure stresses and the
bearing pressure on the piston body considered as a crosshead,
as these appear to be irrelevant. Temperature considerations
determine the proportions of the piston body, and the gudgeon
RUNNING GEAR
169
bearing is made as large as the limited space allows. The
guide pressure between the piston body and the liner works
out at a very moderate figure, and measurements of pistons
after long periods of service fail to disclose any appreciable
wear and justify the conclusion that under normal running
conditions the piston body floats on a ,film of oil. In cases of
seizure the cause of abrasion is usually traced to local distor-
FiG. 127. Type A.
Fig. 127. Type B.
tion, sometimes assisted by the destruction of the oil film by
the presence of viscous deposits.
Water Cooling. — Two systems of conveying the water to
and from the piston are shewn in Fig. 127. In both systems
the aim is to render the success of the scheme independent of
the water tightness of the various joints involved. In type B
inaccuracies of alignment are allowed for by a ball joint at the
foot of the stationary tube.
Pistons for Four Stroke Crosshead Engines. — These are
generally made of not much greater length than is necessary
to accommodate the rings, eight to ten in number. The pro-
170 DIESEL ENGINE DESIGN
vision of an extra number of rings above what is considered
sufficient for a trunk piston may be attributed to : —
(1) The throttling effect lost by discarding the piston skirt.
(2) The lower speed of revolution usually associated with
engines of the crosshead type.
Cooling by means of a blast of air has been used (apparently
successfully) in cylinders of medium size, but water cooling is
now almost universal. Fig. 128 exhibits different forms of
piston having one feature in common, viz., ribbed support of
the crown. There is reason to believe that these ribs are con-
ducive to cracking in uncooled pistons, and therefore the
legitimacy of their use in cooled pistons would appear to
depend on the reliability of the coohng system employed.
Assuming that failure of the cooling system is a contingency
to be reckoned with, there would appear to be some measure
of prudence in proportioning the piston crown to be self-
supporting and transmitting the pressure to the piston rod
via the piston walls. Two systems of water cooling are shewn
in Pig. 129, which requires no description.
Pistons for Two Stroke Crosshead Engines. — ^The existence
of ports in communication with the exhaust pipe at the lower
end of a two stroke cyUnder necessitates the provision of a
skirt or extension of the piston to prevent the uncovering of
these ports when the piston is at the top dead centre. The
skirt usually takes the form of a light drum secured to the
piston by a number of weU-locked studs. It is common
practice to arrange one or two inwardly expanding rings at
the lower end of the cylinder to prevent leakage past the skirt
(see Fig. 130).
If the cyUnder liner is of sufficient length these exhaust rings
may be located in the skirt itself, as in Fig. 131. Such an
arrangement involves a higher engine than that of Fig. 130,
but facilitates conduction oiheat from the piston and inci-
dentally secures a lower mean temperature for the cylinder
liner. The construction of the piston proper is generally
similar to that of a four cycle engine, but it must be borne in
mind that the conditions as to temperature are more severe
than in a four stroke engine of the same size working at the
same mean indicated pressure, so that the remarks of the
preceding article in reference to ribs under the crown apply
with still greater force to two stroke engines. It is significant
RUNNING GEAR
171
FiQ. 128.
Fig. 129.
Outlet
172
DIESEL ENGINE DESIGN
that pistons are now being fitted by one of the leading makers
of large two stroke engines in which the crown is free from
ribs and the skirt is extended to do duty for the piston rod,
as shewn more or less diagrammatically in Fig. 132.
Fig. 130.
Piston Rods.
are : —
Fig. 131.
Fig. 132.
-The forces to which a piston rod is subject
(1) The pressure load of the piston, which attains a maxi-
mum of about 500 lb. per sq. in. of piston area, with
an occasional explosive load of about double this
intensity at the top dead centre.
(2) A load due to the inmia of the piston (including the
water therein) and the rod itself. This also attains its
maximum at the top dead centre, but is opposite in
direction to the pressure load. The maximum in-
tensity of the inertia load seldom exceeds 70 lb. per
sq. in. of piston area in commercial engines.
(3) Friction of the piston rings on the liner. This effect has
its maximum at the top firing centre and acts in the
RUNNING GEAR 173
same direction as the inertia so far as the expansion
stroke is concerned. It has been proved that piston
ring friction absorbs about 5% of the indicated power,
and assuming (as seems probable) that the friction is
at every point proportional to the cylinder pressure,
it appears that the maximum friction is equivalent to
about 10 lb. per sq. in. of piston area.
(4) Fluid friction due to shearing of the lubricating oil film.
The joint effect of these forces amounts then to about 420 lb.
compression at firing dead centre and about 70 lb. tensile at
the end of the exhaust stroke per sq. in. of piston area.
Unfortunately the fatigue stress of steel between limits of
this sort appears not to have been determined yet, but the
value is probably in the neighbourhood of 35,000 lb. per sq. in.
for 30 ton steel. A table of buckling loads for circular mild
steel rods with rounded ends is given below, from which it will
be seen that the question of buckling appears to be irrelevant,
in view of the fact that the ratio of length to diameter of rod
does not usually exceed 15, and is usually less. In this case,
and in cases of connecting rods also, the use of a strut formula,
such as Euler's, in conjunction with a large factor of safety,
would appear to be irrational (see page 184).
L -^D =Length -H Diameter
Buckling stress, Ib./sq. in.
L-HD=Length 4- Diameter
Buckling stress, Ib./sq. in.
In current practice the diameter of the piston rod is made
from 0-23 to 0-3 of the cylinder bore, corresponding to a
maximum compressive stress under normal working conditions.
420 420
of about ^-^-^ = 8000 lb. /sq. in. to ^^:^, = '^^00 Ib./sq. in.
The higher of these figures corresponds to a factor of safety
under fatigue conditions of about 35,000-f8000 = 4-35, which
appears ample. The stress under an exploaon of 1000 lb. per
sq. in. would be =19,000 Ib./sq. in., and the factor of
{O'Zoj
safety against buckling (assuming ==15)=— ^——=2 -35,
which would appear to be sufficient in view of the provision of
a relief valve to prevent the pressure from exceeding greatly
that for which the rod is designed.
5
10
15
20 30
63,000
56,000
45,000
33,000 21,000
40
50
80
100
13,000
9,000
4,000
2,500
174
DIESEL ENGINE DESIGN
The upper end, of the rod usually ends in a circular flange for
carrying the piston to which it is secured by a row of studs
proportioned to the inertia load, of the piston with a very
moderate stress allowance, in order to give a good margin for
dealing with such emergencies as seized
pistons. The lower end is sometimes
secured to the crosshead by means of a
flange, as in Fig. 133, or in the manner
indicated in Fig. 134. In either case
it is reasonable to make the connection
of the same strength as that between
the rod and the piston, assuming that
ample provision has been made for the
contingencies referred to.
Grossheads and Guides. — For
marine engines the slipper guide,
shewn in Fig. 135, is the favourite.
The bearing surface is usually made
equal to the piston area,^ and the
maximum bearing pressure with a con-
necting rod 4-5 cranks long then has
a value of about 55 lb. per sq. in.
The slipper itself is of cast steel, white-
metaUed on ahead and astern faces.
The studs securing the slipper to the
gudgeon block must be adequate to
carry the maximum guide pressure
when running astern. The area of the
gudgeon bearing is based on a bearing
pressure of about 1500 lb. per sq. in.
The ahead guide face is of cast iron
provided with water cooling. The
aa|bern bars are frequently of forged
steel, secured by fitting bolts. The
stress Ta the latter is usually very
moderate, as stiffness is the chief con-
sideration. The proportions given in
Fig. 135 are of course approximate
only and subject to modification to
suit different conditions.
The type of guide block indicated on Fig. 132 is well known
Fig. 133.
Fig. 134.
^ i.e. the sectional area of the cylinder.
RUNNING GEAR
175
in connection with paddle-steamers and locomotives, and
needs no further description here. For land engines double
semicircular guides are sometimes used, particularly when the
cylinder and frame are cast in one piece. In general, the cross-
Width of crosshead face, '75 B.
Depth „ „ 110 B.
Thickness „ plate, 010 B.
Fig. 135.
heads and guides used in Diesel Engine construction differ but
little from those commonly fitted to steam engines and large
gas engines, the most important point of difference being the
gudgeon pin and its bearing, which require to
be liberally dimensioned to withstand the high
maximum pressures to which they are subject.
It is also desirable to provide means to prevent
carbonised oil from the cylinder from reaching
the guide surface.
Guide Pressure Diagrams, — A diagram
shewing approximately the guide pressure at
any crank angle is very simply obtained from
the twisting moment curve in the manner
described below, with reference to Fig. 136.
P= Piston load in lb.
R= Guide reaction in lb.
T= Twisting moment in in. lb.
S= Height of gudgeon pin centre above the
centre of the crank-shaft. j'ie_ i3g_
176 DIESEL ENGINE DESIGN
Now R=-^
But T=P.k
T
Therefore R=-Fr
b
The rule is therefore : Divide the turning moment at any
instant by the distance from the gudgeon pin centre to the
crank-shaft centre, and the result is the guide reaction at the
same instant. It would appear that the guide reaction and the
twisting moment should change sign simultaneously. This is
not quite the case, for the following reason : —
The twisting moment curve contains an inertia element in
which an approximation is obtained by dividing the mass of
the connecting rod in a certain proportion between the revolv-
ing and reciprocating parts. This approximation, though good
so far as vertical forces are concerned, gives very inaccurate
values for horizontal forces. Also the centrifugal effect of the
revolving parts of the rod influences the guide reaction but not
the twisting moment. These discrepancies are of very small
importance with the piston speeds at present obtaining. For
a full discussion of the influence of the connecting rod inertia
forces on the guide reaction, the reader is referred to Dalby's
" Balancing of Engines." A table of values of ^ where " 1 " =
the length of the connecting rod, is given below for various
crank angles, assuming a rod 4-5 cranks long.
Crank angle 0° 20° 40° 60° 80° 100° 120° 140° 160° 180°
Values of |- 1-22 1-21 M6 1-09 1-02 0-94 0-87 0-82 0-79 0-78
Connecting Rods. — The material for connecting rods is
generally Siemens-Martin steel of the same quality as that'
used for the crank-shaft, ^ampings are sometimes used for
small engines, and if large quantities are made at a time this
is an economical way of producing a rod of " H " section, if
machining all over is not considered essential. Cast steel has
been used on the Continent for gas engine connecting rods,
but the author has not met with this practice in Diesel Engine
construction, either British or continental.
Connecting Rod Bodies. — The section of the body or shaft
of the rod is generally circular, or part circular, with flattened
RUNNING GEAR
177
sides. The latter section is slightly lighter for a given strength,
but involves an extra machining operation. For extreme
lightness an " H " section of rod milled from the solid (like a
locomotive rod) would appear to be advantageous. The body
usually tapers gently from the big to the small end, and it will
be shewn later that this practice is a rational one from con-
siderations of strength. It is interesting to note that the
engineers of the North-East Coast, in their recent specification
for standard reciprocating marine steam engines, recommend
parallel rods, and it seems possible that the extra cost of
Fig. 137.
material involved is discounted by the reduced cost of forging
and machining.
Big Ends. — Fig. 137 shews two forms of big end, type A
being the cheapest and that most commonly used. Type B is
the strongest, but suffers from the disadvantage of providing
no facility for adjusting the compression by means of liners.
Returning to type A, the " brasses " are usually of cast steel
lined with white metal. With a stronger section, as shewn in
Fig. 138, cast iron may be used instead of cast steel, with
satisfactory results, but the practice is uncommon. The bolts
are frequently reduced to the core diameter between fitting
lengths, as shewn in Fig. 137 ; but it appears that full diameter
bolts are stronger under the conditions to which they are
178
DIESEL ENGINE DESIGN
subject in trunk piston engines. In addition to tensile stresses
the big end bolts have to resist shearing forces between the two
brasses and also between the crown brass and the palm end of
the rod. Partial relief of this duty is afforded by the following
means, one or more of which are generally used in good
designs : —
(1) Spigoting the two brasses into each other, as in Fig. 139.
This is very rarely done.
(2) Spigoting the crown brass into the palm of the rod,
as in Fig. 137, Type A.
(3) Providing fitting rings, half in the palm and half in the
crown brass at the bolt holes (Fig. 140).
Fio. 138.
Fig. 139.
Fig. 140.
Small Ends. — Various types of small end for trunk engine
rods are shewn in Fig. 141.
With type A the chief difficulty is to find room for bolts of
adequate strength. If a big end of type B, Fig. 137, be fitted,
it becomes necessary to maWe provision at the small end for
adjusting the compression, as in Fig. 141, types B and D.
Type C combines strength and adjustability of the bearing
itself, but makes no provision for altering the compression.
Type E contains a solid bush, which must be replaced when
worn, and is therefore only suitable for small engines, and
adequate section of metal round the bush must be provided to
prevent the hole becoming enlarged (see approximate propor-
RUNNING GEAR
179
tions on Fig. 141). The brasses are usually of phosphor
bronze.
The forked end of a marine connecting rod is shewn in
Fig. 142, on which approximate proportions are noted in
terms of the cylinder bore.
It differs from the similar
member of a marine steam
engine chiefly in the follow-
ing points : — •
(1) The cap brass is not
provided with a steel
keep.
(2) The fork gap is re-
latively narrower,
owing to the piston-
rod nut having to deal
with inertia only.
(3) The brasses are of cast
steel, instead of gun-
metal.
Points of Detail. — The
lubrication of the big and
smaU ends has been referred
to under crank-shafts and
pistons. Where forced lubri-
cation is used an oil-hole is
generally drilled in the rod
to conduct the oil from the
big to the small end in the
case of high speed engines.
External pipes are used for
this purpose in large, slow-
running engines, but are
unsuitable for high speeds
owing to the tendency of
the joints to work loose.
When air compressors, oil
pumps, or other gear are worked by links from the connect-
ing rod the connection to the latter should be made near the
top end, as in Fig. 143, so that the strength of the rod to
resist buckling is not impaired.
180
DIESEL ENGINE DESIGN
Strength of Connecting Rods. — The forces acting on the
connecting rod are : —
(a) The joint effect of
(1) The pressure load on the piston,
(2) The inertia of the piston and crosshead,
(3) The piston-ring friction,
(4) The lubricated friction of piston and crosshead,
all divided by the cosine of the angle of obliquity of the rod.
(b) The longitudinal component of the inertia of the rod
itself.
(c) The transverse component of the inertia of the rod itself.
(d) The friction of the top and bottom end bearings.
Fig. 142.
When considering the compressive stress of the rod on the
expansion stroke one is on the safe side in neglecting item (4).
The tensile forces attain their maximum at the top dead centre
following the exhaust stroke, and the reciprocating parts being
then at their position of minimum speed, item (4) may prob-
ably be neglected with safety.
Item (b) is estimated with sufficient accuracy by the usual
procedure of dividing the mass of the rod between the recipro-
RUNNING GEAR
181
eating and revolving parts in that ratio in which the centre of
gravity of the rod divides its line of centres.
Item (c) gives rise to a bending moment, the maximum value
of which is given approximately by the formula : —
F^ VlOO/
,R.L2
-(1)
26d
Where f =Bending stress.
n= Revolutions per minute.
R=Crank radius in inches.
L=Length of connecting rod in inches.
d=Diameter of rod in inches (mean).
Fig. 143.
Fig. 144.
The sign of the bending moment is such that the latter
always tends to bend the rod outwards to the pide on which
the crank stands. The variation in the magnitude of the stress
over the length of the rod for various positions of the crank
relative to the top dead centre is shewn in Fig. 144 for a rod
five cranks long. The stress varies as sin {d+(f>), where 0=the
crank angle relative to top dead centre and =the angle of
obliquity of the rod, and as Lx— y-, where L is the length of
the rod and x is the distance from the small end to the section
under consideration.
The assumptions made use of in equation (1) are that the
rod is of uniform section, and, as is usually assumed in books
on appUed mechanics and machine design, that the influence
of the rod ends is small.
182
DIESEL ENGINE DESIGN
Item (d) may be estimated, on the assumption that the co-
efficient of friction attains Morin's value of 0-15 for slightly
greasy metal at the top or bottom end or at both ends simul-
taneously. The effect of journal friction is to divert the line
of thrust from the centre line of the rod, and the amount of
this deviation is found by the well-known graphical construc-
tion shewn in Pig. 145, in which the line of thrust is shewn
to be tangential to two very small circles whose radii are
Fig. U5.
Fig. 146.
equal respectively to the radii of the crank-pin and gudgeon-
pin multiplied by the coefficient of friction. The deviation at
any point of the rod of the liift of thrust from the line of centres
will be denoted " d."
The effect of this deviation is to bend the rod into an " S "
shape, as shewn much exaggerated in Fig. 146. This form of
failure is one consistent with the Eulerian theory of pure
buckling, but usually regarded as an improbable solution.
The author has actually seen one instance of a Diesel Engine
rod failing in this way, and the effect was probably due to the
RUNNING GEAR
183
causes indicated above, arising in acute form. To be on the
safe side, it seems advisable to consider two cases : —
(1) Coefficient of friction negligible at the gudgeon-pin and
equal to 0-15 at the crank-pin.
(2) Coefficient of friction negligible at the crank-pin and
equal to 0-15 at the gudgeon-pin.
The weak sections are clearly those (cf. Fig. 149, KK and
SS) where the big and small ends merge into the shaft with a
radius, and incidentally those selected by the draftsman when
Fig. 147.
FiQ. 148.
giving dimensions for the diameters of the rod. The reason-
ableness of making the rod tapered will now be apparent, as
the bending effect is clearly greater at the large end of the rod
on account of the crank-pin being of greater diameteir than the
gudgeon-pin. In all probability these considerations had
nothing to do with the practice, but illustrate a fact that has
a very important bearing on the part of machine design, viz.,
that a construction which appears wrong to the eye of an
individual gifted with a sense of form will usually, on investiga-
tion be found unsound in principle, and vice versa. This
theme is one that might profitably be made the subject-matter
184 DIESEL ENGINE DESIGN
of a less specialised book tan this, but it may be worth
mentioning here that the conscious or unconscious recognition
of this principle appears to be responsible in part for the
ascendancy of continental constructors in certain branches of
mechanical design.
Returning to Fig. 146, it is evident that the deflection at
any point of the rod should, to be strictly accurate, be added
to the deviation of the line of thrust at that point, in order to
find the bending moment, and further, this new bending
moment involves the construction of a revised deflection curve,
and so on. This evidently calls for some form of mathematical
treatment, which with certain approximations can readily be
applied. It will be found, however, that the deflections
involved are small compared with the deviation of the line of
thrust, and whatever error may be incurred can be considered
to be covered by the factor of safety.
On these assumptions, " d " f or the weak sections KK and SS
is given by the following : —
Case (1), Fig. 147 . . d=0-15R<,. t^ (2)
Case (2), Fig. 148 . . d=0-15Rg •?— ^ (3)
And f=pQ+d)__(4)
Where P=The thrust in the rod in lb.
A=Sectional area of rod in sq. in.
Z= Sectional modulus of rod in in. ^
f= Maximum compression stress in lb. per sq. in.
Ro= Radius of crank-pin.
^g= >, „ gudgeon-pin.
It will be noticed that the strut formulae of Euler, Gordon and
Rankine and others have not been utilised above. It appears
to the writer that these formfilse are irrelevant to the case of
Diesel Engine connecting rods, for the following reasons : —
(1) Euler 's formula is based on the calculation of the load
required to produce elastic instability, and with short
rods the stress commonly works out at a higher value
than the ultimate strength.
(2) The Gordon and Rankine formulse are based on experi-
mental values of the buckling stress under static
RUNNING GEAR
185
conditions, and give no indication of the strength
under repetitions of stresses, which are generally only
a fraction of the buckling load.
It seems more rational, therefore, to calculate the maximum
direct stresses as closely as possible and to apply to the approxi-
mately known fatigue stress of steel a
factor of safety of 2-5 to 3, which is
known to be satisfactory in other cases.
In view of occasional abnormal
pressures^of about 1000 lb. per sq. in.,
it is interesting to see what factor of
safety a given rod has for meeting such
contingencies, and the table of buck-
ling stresses given on page 173 may be
used for this purpose.
Example of Stress Calculation for
Connecting Rod. —
Four Stroke Cycle.
Bore of cylinder . . 24 in.
Stroke 30 in.
Revolutions per minute 200
Weight of piston (trunk) 2200 lb.
,, connecting rod
complete . . . 25001b.
Main dimensions of rod, as in Fig. 149,
under : —
(1) Calculation of stress due to thrust
30° after firing centre.
Piston pressure load=0-785 x 24^ x 500 =
Inertia load 30° after dead centre
Fig. 149.
= 226,000 lb.
K
277 X 200V
60
-;
Xl5x
22004-0-35x2500
X(cos30°+icos60°)
386
= 51,000 lb.
Resultant vertical force=226,000-51,000 = 175,000lb.
At 30° after dead centre the obliquity of the rod is 6°.
Connecting rod thrust = 175,000 -^cos 6° = 176,0001b.=P.
At section KK : —
Area = 33 -2 in. "= A
Section modulus =27-0 in. '=Z
186 DIESEL ENGINE DESIGN
Deviation of line of thrust = =0-52=d
75
and the stress f =^176,000 (5^+?^) = 8970 lb. /sq. in.
At section SS : —
Area = 38-5in.2=A
Section modulus =33 -7 in.^=Z
■P, ... .,. ,,, , 0-15x6-25x55 -„_.
Deviation 01 line of thrust = =0'69 in.
75
and the stress f = 1 76,000 (-^ +^^) = 176,000 x -0464
= 81601b./in.2
(2) Calculation of stress due to inertia bending at 30° after
dead centre.
Maximum inertia stress in rod, from equation (1),
2^x15x752 ,.--„ ,. 2
= 26x6-75 =1920 lb. /in.^
From Fig. 145 the fraction of this maximum applying to
postion KK at 30° after dead centre is 0-37,
.'. Inertia bending stress at section KK=0-37xl920
= 710lb./in.2
The fraction applying to section SS at 30°, after dead centre
is 0-49.
.". Inertia bending stress at section SS=0-49xl920
= 940 1b./in.2
(3) Resultant stress at KK=8970-710 = 8260 Ib./in.^
Since the bending actions due to inertia and eccentricity
of thrust are of opposite sign.
Resultant stress at SS = 8160+940=9100 Ib./in.^
since the two bending actions are of the same sign.
(4) Tensile stress at SS at Jaeginning of suction stroke.
Inertia force = (?^)\ 15 x?^5^±|f2i?^^ 1-2
= 63,200 lb.
Stress at SS=63,20o(-i- +|^)=2930 Ib./in.^
The total range of stress is therefore
9100 + 2930 = 12,030 lb. /in.2
RUNNING GEAR 187
The range of stress required to produce fracture of mild steel
by fatigue appears to be about 35,000 lb. /sq. in., so the factor
of safety is about 3.
Calculation on the above lines might with advantage be
made for several different positions of the crank.
It is evident that the results of the calculation depend very
largely on the assumed conditions of journal friction, but it
should be borne in mind that almost any possible combination
of unfavourable conditions is a probable contingency in the
combined lives of a number of similar engines.
Proportions Found in Practice. — In the preceding example
the mean diameter of the rod is approximately 0-28 of the
diameter of the cylinder, a very favourite ratio in practice.
In different designs this ratio varies from 0-26 to 0-30, and it
is a rather curious fact that the two extreme figures are those
which appear to be used by two of the leading makers in their
respective practices.
The maximum and minimum diameters are usually about
5% more and less than the mean.
Connecting Rod Bolts. — In four stroke engines these are
usually proportioned to the maximum inertia load with a
nominal stress of 6000 to 8000 Ib./in.^ based on the inertia
and centrifugal loads divided by the area of two bolts at the
bottom of the threads. With trunk piston engines failure
when it occurs is generally due to piston seizure, to which it
would be difficult to apply definite rules of calculation.
Danger of seizure is largely eliminated by the use of a cross-
head. The strength of connecting rod bolts for four stroke
Diesel Engines forms the subject-matter of a paper by Mr.
P. H. Smith, read by him before the Diesel Users' Association
and containing the results of several years' experience. For
the big end it appears that the bolts seldom fail if made of a
diameter 12 to 13% of the cylinder bore. For the small end, if
bolts are used at all, the only safe rule is to make the bolts as
large as the space available will allow. Mr. Smith also points
out that the bolts, for both big and small ends, are not equally
stressed, as may easily be seen by reference to Fig. 150.
Owing to the deviation of the line of pull from the centre
line of the rod, that bolt (No. 1 in the Fig.) which first passes
the top dead centre at the beginning of the suction stroke is
nearer the line of pull than the other bolt, and consequently
more highly stressed.
188
DIESEL ENGINE DESIGN
If
Then
P= Resultant pull in lb.
S= Centres of bolts in in.
d=0-15 radius of crank-pin.
Pull in bolt No. 1 =
Px - + ci
Px
Pull in bolt No. 2 =
S
With crosshead engines the small end bolts have to carry the
inertia load due to piston, piston rod and crosshead, and also
any frictional forces acting on the piston and crosshead. The
Fig. 150.
latter being more or less indeterminate, it is customary to
allow a nominal stress on these bolts about 30% less than that
allowed for the big end bolts. The bolts or studs connecting
the crosshead to the piston rod and the latter to the piston are
given a large margin of strength for the same reason. Connect-
ing rods for two stroke engines are not as a rule distinguishable
from those for four stroke engines, as the possibility of com-
pression being lost has always to be kept in view.
Indicating Gear. — ^The only satisfactory gear for obtaining
accurate cards consists of a link motion directly connected to
the piston. The usual arrangement is shewn diagrammatically
in Fig. 151 for both trunk and crosshead engines. The condi-
tions for giving an accurate r%roduction of the motion of the
piston are :■ —
(1) The line of the cord to be at right angles to the mean
position of the short arm of the lever.
The long arm of the lever to make equal angles of
swing above and below the horizontal.
The versine of the arc of swing of the drag link to be
negligible in comparison with the stroke of the engine.
(2)
(3)
RUNNING GEAR
189
The mechanical details of indicator gears are hardly of
sufficient interest to require description here, but it may be
well to mention that indicating is of far more importance in the
successful running of a Diesel Engine than in that of a steam
or even a gas engine, and consequently all the more care
should be given to the design of the gear by which the indicat-
ing is accomplished. Makeshift or temporary gear should not
be tolerated, but the same attention paid to lubrication and
bushing of joints, etc., as on other parts of the engine.
Literature. — " Piston Cooling for Diesel Engines." Article
in The Motor Ship and Motor Boat, July 18th, 1918, et seq.
Trunl< Engine. Crosshelad Engine.
Fig. 151.
CHAPTER X
FTJEL OIL SYSTEM
For purposes of description the complete fuel oil system is
conveniently divided, into two parts, the first consisting of
those elements, such as tanks, etc., which are external to the
engine, and the second of those organs of the engine itself
which are directly concerned with the delivery of the fuel to
the working cylinder.
External Fuel Oil System. — Pig. 152 represents diagram-
matically a fuel system for a small Diesel power station and
Fig. 152.
consists essentially of a main storage tank A, a ready-use tank
B, a filter C, and a pump D, fm raising the oil from the storage
tank.
The storage tank is preferably arranged underground, as
close as possible to the railway siding, so that oil can be run
from the railway tank waggon to the storage tank by gravity,
through a hose pipe. Some form of level indicator or a plugged
hole for a sounding rod should be provided. The capacity of
the tank will depend on the size of the station and the local
conditions of supply.
190
FUEL OIL SYSTEM 191
The pump D, by means of which the oil is pumped, to the
ready-use tank, may be of the semi-rotary type, capable of
being worked by one man in the case of small stations ; but
where the daily demand is greater, a motor -driven rotary or
reciprocating pump is generally fitted.
The ready-use tank may have a capacity of say half a day's
run, so that the routine of replenishing it will occur twice daily.
Some form of float indicator should be fitted, so that the level
of oil may be conveniently ascertained from the engine room
floor. Other necessary fittings are an overflow pipe leading to
the storage, or a special drain tank, and a drain valve communi-
cating with the overflow pipe. The tank must be closed at the
top to exclude dirt. The valves in connection with the fuel
system are preferably of the gate or sluice type, as cocks are
liable to leakage and globe valves tend to choke by accumula-
tion of sediment. The tank only requires to be located a few
feet above the level of the filter as the discharge is very small.
The filter usually consists of a cylindrical tank of about
40 gallons capacity, located about two feet above the level of
the cylinder cover and provided with a filter diaphragm at
about a third of the height of the tank from the bottom. The
diaphragm consists of a sheet of felt sandwiched between two
sheets of wire gauze and reinforced by an angle iron ring.
The fuel enters the filter at the bottom, passes through the
diaphragm by virtue of its static head, and is drawn off by the
engine fuel pump at a point a few inches above the diaphragm.
The filter vessel is prevented from being overflooded by a ball
float mechanism which closes the inlet cock when the oil
reaches a predetermined level. The plug of this cock is kept
fairly tight by means of a spring acting on the plug, but slight
leakage is almost inevitable, so it is desirable to mount the
filter on an oil-tight tray provided with a drain. It is very
usual to provide a small reservoir of the same capacity as the
filter arranged alongside the latter for the reception of paraffin,
by means of which the piping leading to the engine, and also
the fuel pumps and fuel valves, etc., may be cleansed from
time to time by running the engine for a few minutes on this
fuel before stopping the engine.
Marine installations follow on similar lines with a few com-
plications. The double bottom is used as a storage tank, and
the fuel is raised to the ready -use tank by motor-driven pumps,
when electric power from auxiliary engines is continuously
192 DIESEL ENGINE DESIGN
available or by means of pumps driven ofi the main engine in
cases where the main engine drives its own auxiliaries. In
either case it is usual to install duplicate pumps to guard against
breakdown. The motion of ships being unfavourable to the
successful operation of float devices, the level of the oil in the
ready-use tank has to be inferred from gauge glasses, test cocks
and the like. For similar reasons, the filters must be totally
enclosed and provided in duplicate with change-over cocks, so
that they may be overhauled at any time. In addition, special
requirements of the Board' of Trade and Lloyd's have usually
to be complied with.
Fuel System on the Engine itself. — The commoner arrange-
ments fall into one of two broad classes : —
(1) Those in which each cylinder has a separate fuel pump
or separate plunger and set of valves to itself. In this
case the oil is delivered direct from the pump to the
injection valve by the most direct route possible.
(2) Those in which one fuel pump plunger supplies the oil
for a pluraUty of cylinders, usually a maximum of four.
In this case the pump dehvers to a fuel main provided
with a branch and distributing valve separate to each
cylinder, whereby the amount of oil delivered to each
cylinder may be equalised while the engine is running.
Engines of six or eight cylinders are divided into two groups
of three or four respectively, so far as the oil system is concerned.
For marine engines the first system is at present the favourite,
and has the advantage that the failure of one pump does not
affect the working of the remaining cylinders. With the
second system a stand-by pump is provided, ready to take
over duty at a moment's notice.
For land engines the two systems appear to be on an equality
as nearly as can be ascertained by the reputations of repre-
sentatives of both classes, but if anything system (1) appears
to be slightly the more popular of the two.
Figs. 153 and 154 illustrate the two systems diagram-
matically. It will be noticed that in Fig. 153 the governor
operates on all the pumps by means of a shaft extending
nearly the whole length of the engine, and as the quality of the
governing is dependent on the freedom from friction of the
governing mechanism it is desirable to mount this shaft on ball
bearings. The expense of providing separate pump bodies
FUEL OIL SYSTEM
193
and, drives is sometimes reduced by grouping the pumps in the
neighbourhood of the governor, even to the extent of driving
all the plungers by a common eccentric.
With the arrangement of Fig. 154 there is only one pump to
regulate, and this renders possible the use of a type of governor
Yue/ Inlet Pipe
Fig. 153.
which is probably unrivalled for sensibility and which will be
described later. The distributors indicated in Fig. 154 are a
special feature of this system and are illustrated to a larger
scale in Fig. 155. A particularly neat arrangement of piping
is obtained by combining the fuel distributor and blast air T-
piece in one fitting.
GovernorW
Fig. 154.
The inclusion of a non-return valve prevents in a great
measure the oil being forced back through the pump by the
blast air pressure in the interval elapsing between the turning
on of the blast air and the attainment by the engine of full
working speed. A non-return valve is sometimes fitted to the
194
DIESEL ENGINE DESIGN
fuel valve itself for the same reason. Vent cocks are provided
on the distributors, and sometimes on the fuel valves, to en-
able the pipes to be primed before starting the engine.
The priming may be effected in various ways : —
(1) By gravity, means being provided for holding the fuel
pump valves off their seats during the process.
(2) By means of an auxiliary hand-operated and spring-
returned plunger on the fuel pump.
(3) By means of the fuel pump plunger itself, where provi-
sion has been made for disconnecting the latter from
its operating eccentric in order to enable it to be oper-
ated by a hand lever provided for the purpose.
To,Fuel Vshe
To Fuel I/aha
FueJ
Brass
Copoer
Via. 156.
The piping in connection with the high pressure fuel system
deserves special attention, on account of the high pressures
used, and the type of uniorrshewn in Fig. 156 is probably the
most satisfactory that has yet been devised both for oil and
high pressure air.
Fuel Pumps. — ^A simple fuel pump for a large marine engine
is shewn in Fig. 157, and is representative of a large class of
pumps for both marine and stationary purposes.
The operation of the pump is almost obvious from the
figure, but may be described briefly as follows : —
FUEL OIL SYSTEM
19S
The eccentric A works the plunger B, which is guided, at C.
E and E are the delivery and suction valves respectively, and
the latter communicates with the suction chamber D, to which
the fuel is led by means of a pipe not shewn in the figure.
M is an auxiliary plunger operated from the crosshead C by
links I, H, K, etc., and whose
function is to keep the suction
valve off its seat for a frac-
tion of the delivery stroke,
depending upon how much
oil is required per stroke.
The duration of this inopera-
tive portion of the stroke is
altered as required by raising
or depressing the point K by
means of an eccentric keyed
to the shaft L, according as
less or more oil is required.
In the case of a governed
engine the shaft L is con-
trolled by the governor. On
marine engines the shaft L
is operated by hand gear,
consisting of levers, rods, etc.
Neglecting the obliquity of
the eccentric rod the' main
plunger describes simple har-
monic motion of amplitude
equal to half the stroke of
the eccentric, and it will be
clear from the drawing that
the auxiliary plunger M will
describe a similar motion
exactly in phase with the
first but of amplitude equal
to stroke of main plunger x
KH
KI
Fio. 157.
When the main plunger is at the bottom of its stroke the
auxiliary plunger is also at the lowest point of its travel, and
the clearance between the top of the auxiliary plunger and the
TCT
suction valve multiplied by the ratio =5™- is equal to the effec-
196
DIESEL ENGINE DESIGN
tive stroke of the pump, that is that portion of the strok<
during which the suction valve is on its seat, as of course r
must be (apart from viscosity effects) for delivery to take place
The quantity of oil delivered per stroke therefore depends oi
a certain clearance between the auxiliary plunger and th(
suction valve, which clearanc(
is readily adjusted by shorten
ing or lengthening the rod LE
when assembling or adjusting
the engine and in the ordinarj
course of running by the eccen
trie at L.
Constructional Details .— Th«
pump body, plunger sleeve anc
guide are of cast iron. Th«
main and auxiliary plungers
the crosshead pin and jointf
in the linkwork are of case
hardened steel. The valves
may be either of steel or cas1
iron. If the latter, then the
suction valve should be fittec
with a hardened steel thimbh
where it makes contact wit!
the auxiliary plunger. Th«
main eccentric and strap ma^
be of cast iron, the lower hal
of strap being white-metaUec
in some cases. It will b(
noticed that no packing is pro
vided for the main plunger
but reliance has been placec
on the fit of the plunger. Wit!
good workmanship the leakage
should not be excessive.* A
Fig. 158.
cast tray is provided to catch drips during working and the
overflow at priming. A light sleeve encircling the auxUiar5
plunger is arranged for operation by external gear so that the
suction valve may be lifted by an emergency governor in cases
of excessive speed, and also by hand in case it is desired to cu1
any individual cylinder out of operation.
Variations of this system, embodying the same principle
* Except with tar oil, for which this arrangement is unsuitable.
FUEL OIL SYSTEM
197
are shewn in Figs. 158 and, 159. The front view of the latter
shews three pumps grouped together, but each worked by its
own eccentric. Fig. 160 shews four plungers being operated
by eccentrics in common. It is evident that with this arrange-
ment the oil delivered to the pulverisers of the various cylinders
wUl have different allowances of time in which to settle before
injection into the cylinder. This appears to have no effect on
the efficiency, but it is usual to space the eccentrics so that oil
is not in process of delivery whilst a fuel valve is open. The
pumps so far illustrated have been driven off the cam-shaft.
That shewn in Fig. 161 is arranged with horizontal plungers
CamshaFt
Hand Pump
Fig. 159.
198
DIESEL ENGINE DESIGN
for driving off a vertical shaft. The auxiliary plunger is driven
by a separate eccentric which on account of the intermediate
lever L requires to be at 180 degrees or thereabouts to the main
eccentric. The suction valve control may in this case be
effected, in one of two ways.
(1) By an eccentric movement of the lever L.
(2) By advancing or retarding the auxiliary eccentric.
" III
ForGoyernor
Control
Fig. 160.
The latter leads to a very neat and efficient arrangement of
governor and fuel pump, to which reference has already been
made. It wiU be immediately obvious that with a given
maximum clearance betw^fci the suction valve and the
auxiliary plunger, an angular movement of the auxiliary
eccentric will have the effect of advancing or retarding the
instant at which the suction valve comes on its seat, and con-
sequently increasing or decreasing the amount of oil delivered
per stroke. This angular movement is effected very simply by
a type of governor which has been well known for a long time,
in steam practice, and which is illustrated in Fig. 170.
FUEL OIL SYSTEM
199
200 DIESEL ENGINE DESIGN
Returning to Fig. 161, the use of this type of fuel pump is
almost entirely confined to land, engines. The provision of an
eccentric mounting for lever L enables the pump to be set in
three different positions, apart from its normal running position,
viz. : —
(1) "Starting." In this position the lever is moved so that
the maximum clearance under the suction valve is
increased about 50%, so that the delivery of oil per
stroke is increased correspondingly.
(2) " Stop." In this position the suction valve is held
continuously off its seat and no oil is delivered.
(3) " Priming." In this position both suction and delivery
valves are held off their seats and the oil has a clear
passage through the pump.
Figs. 163 and 164 wiU make this matter clear without
further explanation.
Details of Fuel Pumps. — ^The bodies are usually of cast iron,
but solid slabs of steel are sometimes used. In designing the
body three considerations should be kept in view : —
(1) The shape to be favourable to sound casting.
(2) As few machining operations as possible to be necessary
apart from those which can be done on a drilling
machine.
(3) The pump chamber and passages to be free from air-
locks.
Owing to the costly precautions necessary to ensure the
plunger and guide being concentric and in line it is convenient
to allow some side play at the point where they join, as in
Fig. 165. Some different forms of plunger packing are shewn
in Fig. 166 and a selection of suction and delivery valves in
Fig. 167. Fig. 168 shiews a hand-operated plunger for priming
purposes.
Calculations for Fuel Pumps. — The process of computing
the capacity of a fuel pump for a proposed engine is most easily
illustrated by an example, as follows : —
B.H.P. of one cylinder (four stroke), 260.
One plunger to each cylinder.
Estimated fuel consumption, 0-4 lb. per B.H.P. hour.
Revolutions per minute, 120.
FUEL OIL SYSTEM
201
202
DIESEL ENGINE DESIGN
Fig. 165.
I
i
m
■h
Fig. 166.
4 4
tIT
mTm
r
Fie. 167.
FUEL OIL SYSTEM 203
Estimated, quantity of fuel per cycle =-^7r — —-=0-0278 lb.
Volume occupied, by 1 lb. of fuel=about 31 cub. in.
Therefore volume of fuel per cycle=0-0278 x 31 =0-86 in. »
Adding 50% to allow for overload, possible increase of fuel
consumption, leakage of plunger, etc. : —
Stroke volume of plunger = l-5x0-86 = l-29 in.^
Which is satisfied by a plunger diameter of | in. x 3 in. stroke.
Fig. 168.
This size of plunger would only be permissible on a marine
engine. If the cylinder belonged to a governed engine the
stroke volume of the fuel pump plunger would need to be
about four times the above figure, as it is found that good
governing at aU loads is only to be obtained by using about
the last quarter of the stroke. This is probably due to the fact
that the quantity of fuel consumed by the engine in a given
time is not proportional to the load but more nearly propor-
tional to the load plus a constant representing the engine
friction. The actual position taken up by the governor and
the effective stroke of the pump plunger at any specified pro-
portion of full load are not easy to determine experimentally
with great accuracy, but the angular positions indicated in
Fig. 169, with reference to the fuel pump eccentric circle, are
those generally used as the basis of calculation.
When one plunger is used to supply several cylinders the
length of effective discharge period is limited by the condition
that the latter should not overlap the injection periods. In
estimating the capacity of a fuel pump driven off a vertical
shaft the speed of the latter must be kept in mind, being usually
the same speed as the engine, and in some cases 50% more.
The valves, hand plungers, etc., are suitable subjects for
distributive standardisation. For example, a suction valve
f in. in diameter would be quite suitable for all sizes of cylinder
(assuming one plunger per cylinder) up to about 20 in. bore
provided that the use of fuels of exceptional viscosity were not
204
DIESEL ENGINE DESIGN
contemplated. For oils like crude Mexican, of the consistency
(when cold) of tar, larger valves are probably advisable. With
the valve arrangements in common use the diameter of the
delivery valve is determined by
that of the head of the suction
valve plus adequate clearance
for the flow of the oil round
the latter.
The general thickness of
metal of the pump body is
usually kept as uniform as
possible to facilitate casting,
and the nominal stress in the
neighbourhood of the pump
chamber based on a blast pres-
sure of 1000 lb. per sq. in. is
about 2500 lb. per sq. in. The
design of a fuel pump affords
ample scope for a draftsman's
skill in many directions, in
which numerical calculation plays a very small part, and the
following suggestions are offered : —
(1) The arrangement generally to be neat and substantial
and presenting an external appearance in keeping
with its surroundings.
(2) The flanges and brackets by which it is secured to the
framework of the engine to be unobtrusive and to
have the appearance of growing as naturally as possible
out of the main body of the casing, so as to convey an
impression of rigidity and equilibrium.
(3) Every detail to be carefully studied, both with regard to
its special function and, also to economy in manu-
facture, efficiency always having precedence over
economy. In particular, case-hardoning and bushing
of parts subject to wear must not be stinted, and
provision should be made for lubrication of aU moving
parts.
(4) Valves and other internal mechanism to be easily
accessible for inspection and overhaul.
(5) Arrangements to be made to catch all drips, both of fuel
and lubricating oil, avoiding the use of trumpery
sheet-iron guards and the like.
FUEL OIL SYSTEM
205
Fig. 170.
206 DIESEL ENGINE DESIGN
Many of the above principles apply of course to the design
of any part of any high-grade machine, and they are mentioned
here because the matter on hand provides an excellent oppor-
tunity of emphasizing them in a particular case, in which the
subject is singularly free from the complications arising from
calculations. When the discussion is transferred to some large
part of a machine, in which the stresses are approximately
determinate and the scope of the design appears to be limited
by adjacent parts, it becomes increasingly difficult to reconcile
the ideals of high-class design with the requirements of effi-
ciency and economy and the skill of a designer may be gauged
by the extent to which this difficulty is overcome. From this
point of view no part of a design can be said to be finally deter-
mined until the whole design is complete, as there is always
the possibility that a design for a certain part, perfect in itself,
may require to be modified subsequently on account of its
relationship, perhaps remote, to some other part as yet un-
determined.
Governors. — It is not proposed to deal here with governor
design generally, as that is a subject for a specialist in this
particular department of mechanical design, but only to
illustrate the application of governors to stationary Diesel
Engines by means of a few examples, and to give the main
lines of calculation for the type of governor shewn in Fig. 170,
which is a type not usually standardised by governor specialists.
The action of the weights in causing rotation of the central sleeve
will be immediately obvious from the figure. The amount of
this rotation between the limits of no load and full load should
correspond with the angle ~ 60° of Fig. 169, but as a safety pre-
caution it is advisable to give the governor sufficient range to
give a complete cut-out, and the sleeve should therefore be
free to describe an angle of about 70°. The first stage in the
design of the governor is to rough out a drawing similar to
Fig. 170, fulfilling all the reauirements as to space, accessi-
bility, etc., and in which the aloove angular rotation is secured.
As regards the size of the governor, it is generally wise to avail
oneself of all the space obtainable. The next step is to find the
mass and centre of gravity of the weights and the positions of
the latter in the extreme in-and-out positions. A diagram
similar to Fig. 171 should now be constructed, in which the
abscissae are distances of the weight from the centre of the
shaft in inches and the ordinates are the centrifugal forces at
FUEL OIL SYSTEM
207
these distances at no load speed and full load speed respectively.
Point " p " corresponds to "no load " speed and " no load "
distance from centre, and point " Q " the same quantities for
full load. The line PQ then determines the properties which
the controlling spring would have to possess if it were con-
nected to the weight at its centre of gravity. These properties
are as follows : —
(1) The initial tension, when the weights are full in, is equal
to the centrifugal force corresponding to the point
"Q."
(2) The weight of the spring per inch extension is equal to
the slope of the line PQ, that is the amount in lb. by
which the ordinate increases as the abscissa increases
by one inch.
Actually the spring is attached to the weight at a point
nearer to the fulcrum than the centre of gravity, and both the
initial tension and the rate as found thus require to be increased
in the ratio t-, where : —
k=Distance of the weight fulcrum from the line joiniag
the C.G. of the weight to the centre of the governor.
1 =Distance of the weight fulcrum from the line of action
of the spring.
Radius of C.G. oF Goifernor Weight
Fig. 171.
This very simple construction, repeated as often as may be
necessary in the process of trial and error, contains all the
dynamical calculation required to ensure sensibility and
stabihty, but it is advisable to provide adjustments for spring
tension, in the manner shewn in the figure, to allow for un-
208
DIESEL ENGINE DESIGN
avoidable errors and routine adjustment on the test-bed.
Strictly speaking, the diagram shewn in Fig. 171 should be
corrected to allow for the versed sine of the arcs described by
the points of suspension, and so on ; but these are practically
negligible. Other types of governor are designed on similar
lines, but are usually complicated by Unk mechanism, of which
the variations of configuration are not negligible.
Fig. 172.
Fig. 173.
Variation of speed during the running of the engine is readily
seciu-ed by transferring a ^rt of the controlling force to an
auxiliary spring, the tension of which can be varied by
mechanism provided for the purpose, as shewn in Fig. 172,
or by varying the tension of the main spring itself, as in
Fig. 173.
Some points to be observed in governor design are : —
(1) Weights as heavy as possible, to give power and con-
sequently render the effects of friction negligible.
FUEL OIL SYSTEM
209
(2) Springs to be readily adjustable.
(3) Small pin-links, etc., to be as substantial as conveniently
possible.
(4) Friction to be reduced to a minimum by ball-bearings.
(5) Joints other than ball-bearings to be bushed and pro-
vided with weU-hardened pins.
(6) Lubrication, both as regards supply of lubricant to the
working parts and systematic disposal of the surplus,
to be considered carefully.
The general disposition of the governor and fuel pump with
respect to the framework of the engine is shewn in Figs. 174,
Fio. 174.
Fig. 175.
175 and 176, in three cases. In Fig. 174 the governor is of the
angular movement type described above and illustrated in
Fig. 170, driven off the vertical shaft from which the cam-shaft
receives its motion. The fuel pump is of the horizontal plunger
type, receiving its motion from eccentrics mounted on the
same vertical shaft. The fuel pump body is supported by a
facing on the lower side of the case containing the upper spiral
gears.
In the arrangement shewn in Fig. 175 the fuel pump is
attached to the cylinder jacket, but in other respects the
details are similar to those of Fig. 174.
The governor shewn in Fig. 176 is of the more usual type
characterised by a sleeve which is mounted on a feather and
which rises as the engine's speed increases. The pump is of the
vertical multi-plunger type, and regulation is effected by
210
DIESEL ENGINE DESIGN
rotation of an eccentric shaft, on which are hinged the levers
which operate the auxiliary plungers.
The arrangements described, briefly cover the bulk of the
Fig. 176.
fuel pump and governor mechanisms found in practice, but
mention must be made of some modern refinements which are
coming increasingly to the fore.
(1) Control of fuel valve opening. At light loads the dura-
tion of opening of the fuel valve is greater than
necessary if uncontrolled and the instant of opening
which is most favourable for full load running is
inclined to be late for light load running. At least one
firm has attacked this problem of governor control of
the fuel valve operating mechanism.
(2) Blast pressure control. This question is closely allied
with (1), as a shortened opening period would lead to
increase of the blast pressure if the latter were not
corrected. Apart from this the blast pressure has in
any case to be altered in accordance with the load
(unless cylinders are cut out of operation) if good
combustion is to be secured at all loads, including no
load. The blast pressure is placed under the control
of the governor by means of a throttle slide on the
compressor suction.
(3) Pilot ignition. This refers to cases where exceptionally
refractory oils are being used which require for their
FUEL OIL SYSTEM
211
combustion a preliminary charge of a lighter oil, such
as Texas oil, which is deposited in the pulveriser in
advance of the main charge by a small auxiliary pump
provided for the purpose. The necessity for this
device appears likely to be obviated by improvements
in fuel valves and the flame plates in particular.
Fuel Injection Valves. — It now remains to deal with the
valve by means of which the fuel is injected into the combus-
tion space and to which oil is delivered by the fuel pump for
this purpose. These may be broadly classified as the open and
closed types respectively, and as the former form a relatively
small class at present it is convenient to dispose of them first.
Open Type Fuel Valves. — Fig. 177 is a diagrammatic view
of such a valve, omitting all detail not required to illustrate
Fig. 177.
the bare principle. Oil is delivered to the space A past the non-
return valve B by means of the fuel pump, and this type of fuel
valve derives its name from the fact that the space A is in
constant communication with the interior of the cylinder. It
is to be noticed that the fuel pump is not required to deliver
against the pressure of the blast air as the latter is restrained
by valve C. The latter is opened by appropriate gear at the
predetermined instant for injection and carries with it the fuel
oil contained in the space A. The action appears to be highly
eflftcient in pulverising effect and excellent fuel consumptions
have been reported for engines in which these valves have been
fitted. This type of fuel valve appears to have been devised in
the first instance for use in horizontal engines in which it was
anticipated that the more usual type of fuel valve would be at
a disadvantage. A valve working on a somewhat similar
principle has been tried, from all accounts successfully, on
212
DIESEL ENGINE DESIGN
vertical engines, but has not yet, to the author's knowledge,
become a standardised fitting.
Closed Type of Fuel
Valves .— In this type com-
munication between the
combustion space and the
interior of the fuel valve
only exists during the in-
jection period, when the
flow is always in the same
direction, apart from such
derangements as stuck
valves or failure of the
blast pressure.
Fig. 178 shews what
may not improperly be
called the Augsburg type
of fuel valve. Apart from
the cast-iron body, the
construction of which is
sufficiently illustrated by
the drawing, the principal
parts are : —
(1) Needle valve A.
(2) Spring B.
(3) Stuffing box C.
(4) Pulveriser D.
(5) Elame plate E.
The needle valve is
usually made of special
steel, case - hardened in
way of the stuffing box
to prevent cutting by the
packing. Accurate align-
ment of aU parts of the
needle is essential and
readily secured by grind-
ing between fixed centres.
The lower part of the
needle is preferably re-
duced in diameter by a
FUEL OIL SYSTEM
213
few thousandths, as a certain temperature gradient exists
between the needle and the pulveriser tube which may lead to
seizure if sufficient clearance is not allowed. The tip generally
has an angle of about 40 degrees. The needle spring, in addi-
tion to returning the needle to its seat against the pressure of
the blast air, has to deal with the friction of the stuffing
box, and may be figured out on the basis of a pressure of
1500 lb. per sq. in. over the sectional area of the needle at
the stuffing box. The latter is usually provided with a
screwed gland.
Extra hole For Pi I
Ignition charge
.',:_ 00 1 to Q-0I3B
Fig. 179.
The pulveriser tube is held on its seat by a stiff spring, and
serves the double purpose of affording some support to the
needle and retaining in their relative positions the rings and
the cone which play an important part in pulverising the fuel.
It will be clear from the figure that the pulveriser is sur-
rounded by blast air, which enters at F. The fuel is introduced
by means of a narrow hole G, at a point H immediately above
the top ring. If the point H is located too high the oil fails to
distribute itself evenly round the pulveriser rings and in-
efficient combustion results.
Fig. 179 shews the injection end of the pulveriser, together
214
DIESEL ENGINE DESIGN
with the flame-plate and nut, to a larger scale. The details
shewn are those in most common use, but are subject to
variation in the practices of different manufacturers. The
proportions shewn are roughly indicative of good practice,
but it must be admitted that the rule of linear proportionality
does not appear to be rational in this case. Experience in this
matter discloses two facts : —
(1) That for a given engine there is a certain minimum
diameter of pulveriser ring, below which results are
not satisfactory (about 9% of the cyhnder bore).
(2) That as cylinders are increased in size it becomes in-
creasingly difficult to obtain a high M.I.P.
These suggest the following hypothesis : —
That the best results are to be obtained when the depth of
oil in the pulveriser before injection is a certain amount, and
the same for cylinders of all sizes. If this is true, then the ared,
of the pulveriser ring should be in proportion to the cylinder
volume. This would lead to the diameter of pulveriser rings
being made proportional to the cylinder bore raised to the
power of 1'5. Such a rule has not been adopted, and would
probably lead to inconveniently large valves in the larger sizes
of engines, but the question would appear to offer some induce-
ment to research. A very large number of different types of
pulveriser are in use, and have been described in the technical
press ; but it still remains to be proved
that they are more efficient than the
common variety shewn in Pig. 179. A
neat form of pulveriser tube, which dis-
penses with the long narrow hole drilled
in the fuel valve casting, is shewn in
Eig. 180, from which it will be seen that
the oil is led to an annular space A at
the top of the tube, whence it flows down-
wards to vke pulveriser rings via a number
of grooves in the surface of the pulveriser
tube. Holes B are provided to give passage
for the blast air.
Swedish Type of Fuel Valve.— Eig. 181
shews the construction of this type of
valve, which has also been widely adopted
and which is characterised by the fact that
Fio. 180. "^
FUEL OIL SYSTEM
215
216
DIESEL ENGINE DESIGN
M~
the needle is completely enclosed, within the casing and is
subject on all sides except the extreme tip to the pressure of
the blast air. On this account the spring A does not require
to be as strong as that of an Augsburg type of valve of the
same size. The needle is lifted in working by the lever B
attached to a cross-shaft C, the end of which penetrates the
casing through a stuffing box D. The mechanical means by
which end thrust on the cross-shaft and bend-
ing actions on the overhung end, due to the
pressure on the external lever, are dealt with,
will be clear from the figure without further
explanation. The use of this type of valve
appears to be limited at present to those
designs in which the requirements of other
parts of the valve gear necessitate the fuel
valve operating lever being arranged off the
centre line of the cylinder cover.
Burmeister Fuel Valve. — ^The construction
of this valve is shewn diagrammatically in
Fig. 182, and its outstanding features are the
use of a mushroom valve, the extreme sim-
plicity of the whole arrangement, and the fact
that the valve is opened by a downward
movement. The latter is a particularly valu-
able feature as it secures uniformity of valve
gear and ease of withdrawal.
The four classes of fuel valve described
above include as members practically all the
fuel valves in use on Diesel Engines at present.
Each type has its advantages, but no one of
them can be said to hold the field. Something
similar might be said for the enormous variety
of pulverisers patented and in actual use. It
seems doubtful if any of these can claim out-
standing efficiency. When *lot injection of a less refractory
oil is used to facilitate the use of tar oil as fuel an additional
hole has to be drilled in the fuel valve, as shewn dotted in
Fig. 179. The question of burning tar oil is still in the experi-
mental stage in this country, but the results so far obtained
hold out hopes that it will be possible to dispense with pilot
ignition in favour of special arrangements of a simpler char-
FlG. 182.
aoter in connection with the fuel valve details.
FUEL OIL SYSTEM
217
Some of the arrangements by means of which fuel valves
are operated are shewn in Figs. 183, 184, 185 and, 186. The
long lever, which is a feature of all these schemes, is usually of
cast steel, and should be of stiff construction. The fulcrum on
which the lever hinges is common to the levers which operate
Fig. 183.
Fig. 184.
the other valves, viz. air and exhaust valves in the case of four
stroke engines and scavenge valves in the case of two stroke
engines, and starting valves in both cases. With land engines,
and many marine engines, it is usual to mount the fuel and
starting valve levers on eccentric bushes mounted on the
Fig. 185.
Fig. 186.
fulcrum shaft at such angles that the operation of putting the
starting valve into gear automatically puts the fuel valve out
of gear and vice versa. This is considered in detail in
Chapter XI.
The use of the needle type of fuel valve in conjunction with
a single lever necessitates the latter being so disposed that its
218 DIESEL ENGINE DESIGN
roller is rendered, more or less inaccessible by the cam-shaft,
particularly if the latter runs in a trough (see Fig. 184). The
difficulty may be got over by providing a small intermediate
lever, as shewn in Fig. 185, to reverse the direction of motion.
In spite of the objections which have been raised against this
arrangement it appears to be satisfactory in practice.
Design of Fuel Valves. — An approximate rule for the
internal diameter of the body has already been given, being
the same as the diameter of the pulveriser rings. The thick-
ness of the walls (cast iron) may be from a third in large valves
to a half in the case of small valves of the internal diameter.
If the valve is of the Swedish type this thickness will be approxi-
mately constant throughout the body of the valve, except in
the neighbourhood of flanges, etc. If of the Augsburg type,
those parts of the body not subject to pressure may be a little
thinner. In all cases a good rigid job should be aimed at, as
lack of alignment leads to sticking of the valve. The pulveriser
tube is made of steel and the details, such as rings and cones, of
steel or cast iron. The Swedish type of valve requires special
care to be devoted to the design of the cross-shaft and its fit-
tings, in order to obtain freedom under load, adequate bearing
surface and accessibility of the stuffing box. As regards the
valve as a whole, the designer should aim at shapely solidity
and avoid flimsiness of detail.
With the Augsburg type of valve (Fig. 178) the load necessary
to lift the needle is the spring load less the product of the blast
pressure and the area of the needle at the stuffing box (approx.)
plus the gland friction. With the Swedish type (Fig. 181) the
load may be taken as approximately equal to the spring
pressure plus the product of the blast pressure and the area of
the needle at its seat. This load evidently induces bending
and twisting actions, which the cross-shaft should be pro-
portioned to carry with a low stress. The weakest section is
generally at the reduced diameter to which the external lever
is keyed. The key itself shwild be amply proportioned, and is
preferably made of tool steel. The ball thrust must be pro-
portioned to the load obtained by the product of the maximum
blast pressure into the sectional area of the cross-shaft at the
stuffing box. The flame plate is of nickel steel and the dia-
meter of the hole is usually about 1% of the cylinder bore, but
the best size for any particular case must be found by experi-
ment. The flame plate nut may be of steel or bronze secured
FUEL OIL SYSTEM 219
to the fuel valve body by a fine thread and provided with flats
to accommodate a spanner.
The main points in the design of fuel valves may be sum-
marised as follows : —
(1) Rigidity and alignment of casing.
(2) Alignment of the needle and its guide.
(3) Freedom of all working parts.
(4) Sturdy proportions for all small details.
Literature. — Renolach, N. 0., " Tar Oils as Fuel for Diesel
Engines." — Internal Gombusiion Engineering, July 15th, 1914,
et seq.
Porter, G., "Tar Oil Fuel and Diesel Engines." — Diesel
Engine Users' Assoc, May 24th, 1917.
Smith, P. H., "Two Essential Conditions for Burning Tar
Oil in Diesel Engines." — ^Diesel Engine Users' Assoc, May
16th, 1918.
CHAPTER XI
AIR AND EXHAUST SYSTEM
Four Stroke Engines. — So far as four stroke engines are
concerned the parts included in this system are : —
The air suction pipe.
Air suction valve.
Exhaust valve.
Exhaust piping.
Silencer.
In the neighbourhood of cement works, or other sources of grit,
a suction filter is sometimes added, with a view to preventing
foreign matter from reaching the interior of the cylinder along
with the indrawn charges of air.
Air Suction Pipes. — The usual form of
suction pipe for both land and marine engines
" is shewn in Fig. 187, and is intended to serve
the double purpose of muffling the sound
caused by the rush of air past the inlet valve
on the suction stroke and also to prevent in
some measure the ingress of dust. To be of
real service in either capacity the slots should
not exceed about 40 /lOOO of an inch in width,
and in use must be kept clear of dirt, other-
wise the volumetric efficiency, and consequently
the power of the engine, will be seriously
impaired. To be reasonably effective as a
silencer th^ slots should cease within a foot
or so of the point where the pipe joins the
cylinder cover. To prevent throtthng the
aggregate area through the slots is usually
made about 1-5 to twice the clear area of the
pipe. Assuming the slots are spaced f " apart,
the slotted length of the pipe works out to
about 3i to 4 J times the bore. In very
220
Fio. 187.
AIR AND EXHAUST SYSTEM
221
expert foundries these slots are formed in the casting (cast
iron). Where the slots have to be milled aluminium is a
convenient but expensive material for this purpose. Welded
tubes of sheet iron are sometimes used, but are liable to become
dented and unsightly.
An efficient and durable arrangement, shewn in Fig. 188,
consists of a common collecting pipe in communication with
all the cylinders and ending in a trumpet-shaped piece which
is very effective in muffling the sounds of suction. The trumpet
fitting is similar to a cornet mute, and consists of internal and
external members so arranged that the space between them
has a sectional area increasing outwards whilst the distance
between the two members diminishes.
The noise question is most effectively disposed of by carry-
-f
111
Fig. 188.
ing the suction pipe outside the building in the case of a land
engine or on deck in the case of a marine engine. If an air
filter be added the arrangement is ideal but expensive.
Where a separate suction pipe, as in Fig. 187, is fitted to each
cylinder the bore is usually made equal to that of the suction
valve or slightly larger. Where a common suction pipe is
provided the bore may be anjrthing up to about double this
size, according to the number of cylinders to be supplied and
the length of the pipe.
Suction Valves. — Atypical suction valve is shewn in Fig. 189.
The valve proper may be of carbon or nickel steel, in the form
of a drop forging. The upper end is guided by a little piston,
which also takes the thrust of the spring. In the example
shewn the guide piston is of chilled cast iron extended in the
form of a nut and provided with a cup-shaped cavity to accom-
modate the operating tappet. The casing being in two parts,
it is necessary to finish each part on a mandril to ensure align-
ment in position. It is not necessary at this point to describe
222
DIESEL ENGINE DESIGN
variations in the details of suction valves, as the latter are
usually made similar to the exhaust valves (which wiU be
described in detail later) except for such special features as are
necessary to deal with the heat effects to which the latter are
subject.
FiQ. 189.
Dimensions of Air Suction Valves. — ^The requisite diameter
for an air suction valve may be regarded as determined by the
mean vacuum allowable on the suction stroke. Taking this to
AIR AND EXHAUST SYSTEM 223
be 0-6 lb. below atmospheric pressure, the theoretical mean
velocity is found from Fig. 13 to be about 280 feet per second,
and taking the mean coefficient of discharge to be 0-70 this
gives a mean apparent velocity of 195 feet per second. Now as
regards the mean opening area of the valve, if we assume a
harmonic cam opening and closing exactly at the upper and
lower dead centres respectively, then the mean opening would
be just half the maximum opening if the maximum Hft is made
= J of the value diameter. Actually the valve is always
arranged to open before top centre and close after top dead
centre, so that the mean area is usually more like 0-6 of the
maximum area.
Adopting this figure, we may write : —
V^ ^P^0-65x0-785d2
Where Vj= Apparent mean velocity of air in feet per second.
Vp=Mean piston speed in feet /seconds.
B =Bore of cylinder.
d= Diameter of suction valve.
From which,
g-^I-54-^-^
p_
127
Values of = calculated from this formula for various piston
X)
speeds are given below and agree well with average practice.
Piston speed in
ft. permin. . 500 600 700 800 900 1000 1100 1200
Ratio d-f-B . -257 -281 -303 -324 -343 -362 -380 -397
In the best practice the maximum lifts of the air and exhaust
valves are frequently made as much as 0-35 of the valve dia-
meter, as although the extra lift does not increase the maximum
available area the mean area is increased and the opening at
the dead centre is augmented without adopting an unduly long
opening period or an awkward shape of cam.
The air and exhaust valves are usually operated by the lever
arrangement shewn in Fig. 186 with reference to fuel valves.
Exhaust Valves. — In some small engines the exhaust valves
are similar to and interchangeable with the air suction valves,
but with medium and large size engines, owing to the larger
dimensions of the exhaust valves, and in consequence their
slower rate of cooling by conduction to the surrounding media,
224
DIESEL ENGINE DESIGN
special arrangements have to be made to conduct heat away
from the valve seat which would otherwise become pitted and
grooved in a short time.
With cyhnders up to about 16 in. bore the trouble can be
reduced to reasonable proportions by providing the valves
with cast iron heads, which are less liable to become pitted
than those of nickel steel. Ventilation of the casing by means
of perforations is also found useful. Some types of cast-iron
exhaust valve heads which have proved successful in practice
are shewn in Fig. 190. With cylinders much in excess of
16 in. bore some form of water-cooled casing is desirable, and
the simple arrangement shewn in Fig. 191 has proved most
Cast Iron
Cast Iron
f Iron
Fig. 190.
Fio. 191.
AIR AND EXHAUST SYSTEM
225
effective with the very largest cylinders without resorting to
the expedient of coohng the valve itself by direct means. It
appears that the flow of heat from the seating to the water-
jacket is suflBicient to keep the temperature of the former at a
suitable low value so long as the mean indicated pressure does
not exceed a reasonable figure dictated by other considerations.
The valve spindle is liable to become stuck by carbonaceous
deposit, and should therefore be about 20/1000 slack in the
casing. Furthermore, arrangements should be made to feed a
little paraffin to this point from time to time.
The upper or guide end requires some care in detailing,
Fig. 192.
particularly in high speed engines, to prevent the nut from
slacking back or the threads becoming stripped. The arrange-
ment shown in Fig. 189 is simple and effective ; other arrange-
ments are shown in Fig. 192 for slow speed engines.
Valve Casings. — Two types of valve casings have been
illustrated in Figs. 189 and 191, and further modifications are
shown in Fig. 193. The valve seating is made in a separate
piece, which is readily replaceable when worn by an inter-
changeable spare. After repeated regrindings a ridge is formed
which has to be turned off, a process more conveniently carried
out on a light ring than on a whole casing. If no loose seat is
provided it becomes necessary in time to turn down the under
surface of the casing flange, in order to bring the valve seat
226
DIESEL ENGINE DESIGN
flush with the under side of the cover. The port in the easing
provided for the flow of exhaust gas should accurately coincide
with the corresponding port in the cover, and it is usual to
provide an equal and opposite dummy port to preserve the
symmetry of the casting and diminish chances of distortion
due to expansion or growth. The
depth of the metal under the ports
should not be reduced to too fine a
limit, otherwise there is danger of
leakage on the compression and ex-
pansion strokes, due to deflection or
even cracking of the metal at this
point. The number of studs (two to
four) used to secure the casing to the
cylinder cover is purely a matter of
convenience in arrangmg the gear on
the cover or cylinder head. Two studs
properly proportioned are sufficient
for the largest valves, given an
adequate depth of flange and good
connection of the latter to the body.
Exhaust Lifting Devices. — Some
form of exhaust lift for breaking com-
pression is usually fitted to both land
and marine engines. With the former
the use of such a device on shutting
down the engine obviates the ten-
dency of the engine to swing in the
reverse direction to that for which it
was designed, just before stopping,
and also facilitates turning the engine
by hand. With marine engines some
such device should come into opera-
tion automatically on reversing to
present the possibility of compressing
an unexhausted charge when motion is
begun in the reverse direction. An expansion stroke executed
in the ahead direction becomes a compression stroke in the
astern direction and vice versa. Two hand devices are shown
in Fig. 194. Tjrpe A consists of a substantial steel lever pro-
vided with a slotted end which normally keeps it in a vertical
position clear of the gear. To bring the lever into operation it
Fig. 193.
AIR AKD EXHAUST SYSTEM
227
is lifted to the extent of the slot and allowed to fall forwards.
A projection on the lever then slips under an extension of the
roller-pin on the first occasion of the valve being lifted and pre-
vents its return. Type B consists of a link and screw, by means
of which the valve may be depressed during the running of the
engine and which normally lies alongside the casing. Fig. 195
shows three arrangements suitable for marine engines. In type A
the valve is depressed by a pneumatic cylinder arranged above
Fig. 194.
and in line with the valve. Type B consists of a series of cams
(one cam over each exhaust valve lever) mounted on a shaft
running from end to end of the engine. A turning movement
of this shaft simultaneously depresses all the exhaust valves.
Type C is appropriate to those engines in which the valve levers
are operated by push rods. The first movement in reversing
consists of swinging the lower end of the push rods out of the
range of operation of the cams. By a scheme of linkwork,
which is obvious from the illustration, the same movement
introduces a lever with an incUned face under a roller provided
for this purpose attached to the valve lever, with the result
that the valve is held off its seat until the push-rods regain their
normal working positions.
Proportions of Exhaust Valves and Casings. — The diameter
of the exhaust valve is almost invariably made the same as
that of the air valve, in order to simplify manufacture and give
228
[DIESEL ENGINE DESIGN
a symmetrical arrangement of valve gear. On the exhaust
stroke the area through the valve is more than sufficient, but
a fairly early opening (40 to 50 degrees before bottom dead
centre) is necessary to effect a rapid fall of pressure. The
dimensions of the various parts of the valve and casing are
with a few exceptions matters of experience only, and the
approximate proportions given below with reference to Fig. 196
are representative of average practice. The figures are ex-
pressed in terms of the valve diameter as unit.
■ B"
Fig. 195.
The larger figures refer generally to the smaller sizes of valve.
The spring, which wiU be considered later, has usually between
•twelve and twenty-two turns. A large number of turns reduces
the range of variation of stress, and consequently increases
resistance to fatigue.
On the other hand, a spring with a small number of turns
has less tendency to buckle^
The seating of the casii^ in the cylinder cover is almost
always made square now instead of conical as formerly. The
width of seating need not exceed a quarter of an inch in the
largest sizes. The seating of the valve in the casing is usually
made at an angle of 45 or 30 degrees, and the width of the
seating varies greatly in different designs. Narrow seatings
about an eighth of an inch in width appear to be least subject
to pitting, and seatings as narrow as l/25th of an inch have
AIR AND EXHAUST SYSTEM
229
been used successfully. For large marine engines seats about
f inch wide seem to be preferred. The spindle clearance may
be about 20/1000 and the guide piston clearance about 10/1000
in all sizes.
The size of the holding-down studs may be found from the
standard table (page 130), the total load being based on a
pressure of 500 lb. per sq. in. over the least area of the casing
where it makes joint with the cylinder cover.
The casing lugs shoiold be amply proportioned, particularly
in cases where the castings are not above average quality, and
they should have a good hold on the cylindrical part of the
casing, reinforced if necessary by internal ribbing. It is not
■h 12 to?2 turns
a
6
c
e
/
h
015
0-20
046
007
046
006
to
to
to
to
to
to
018
0-24
056
OIT
0-64
008
Fig. 19
unknown for these lugs to break off in tightening up the studs,
so it is as well to err on the strong side.
Exhaust Valve Springs. — Apart from the weight of the
valve in itself there are causes tending to open the exhaust
valve when it should be shut, viz. : —
(1) The vacuum on the suction stroke.
(2) With four cylinder engines especially, the first rush of
exhaust from the neighbouring cylinder if the latter
discharges into the same collecting pipe.
In addition, the inertia of the valve and any levers, rods,
etc., in connection therewith, tend to make the latter lose
contact with the operating cam in the neighbourhood of
maximum lift. These various influences are overcome by
fitting a spring, the normal load of which is equivalent to a
230
DIESEL ENGINE DESIGN
pressure of about 10 lb. per sq. in. of valve area in the case of
slow speed engines, and anything up to about 20 lb. per sq. in.
or more in the case of high speed engines. A method of comput-
ing the inertia effect will be dealt with in some detail as there
is a tendency for higher speeds to be used in practice, and the
principles involved have a wide application in the design of
high speed machinery generally.
Inertia Effect of Valves. — Consider the simple arrangement
shown in Fig. 197, consisting of a valve directly operated by a
cam without intermediate members. Let the weight of the
valve guide and roller, etc., be " W" lb.
The effect of the inertia of the spring
may be allowed for by adding one-third
of its weight to that of the other parts.
Let " X " be the distance in inches of the
valve from its seat at any instant. If the
shape of the cam and the speed of the
cam-shaft be known it is possible to ex-
press " X " in terms of the time " t " in
seconds counted from the instant at
which the valve begins to lift, by means
either of an equation or a graph exhibiting
the lift on a time base. If this equation
(equation of motion) is available, then
one differentiation with respect to "f'
gives an expression for the velocity
denoted by " x," and a second differentia-
tion gives the acceleration denoted by
" X." If the relation between x and t is
given by means of a graph, then the
differentiation may be done by one or other of the graphical
methods explained in books on practical mathematics. In
either case, x being measured from the valve seat outwards,
positive values of x denote inertia effects tending to press the
roller against the cam, and negative values of x denote inertia
effects tending to cause the roller to lose contact with the cam.
Here we are only concerned with the negative values of x. If
X denotes the resultant force on the valve, neglecting all
Fio. 197.
effects except those due to the inertia,
W
Then, X = — .XntrA-g. (1)
g=386in./sec.«
AIR AND EXHAUST SYSTEM 231
Equation (1) gives the minimum value of the spring pressure
to prevent the roller jumping due to inertia.
The value of x (max.) is very easily calculated in one case,
viz., when the cam is so designed that the valve describes
simple harmonic motion ; that is, when the graph of x and t
is a sine curve. In this case it is convenient to measure x from
the position of mid-Uft positive outwards and negative inwards.
The equation of inotion is then : —
x=A.sin pt — — (2)
where A = Half the maximum lift in inches,
and p is a constant such that pT=27r,
where T is the whole period of opening in seconds. If the
exhaust valve is open for 240 crank-shaft degrees, then : —
_ 60^240 , 27r.nx360 „ _„
^=n ^360' ^""^ P=^60^^240-='-^''^— ('^
" n " being the number of revolutions of the engine per minute.
From (2) x= — Ap^.sin pt.
And X(jj^x)= - Ap"./
W
Substituting in ( 1 ) X = — ^ . A.p ^ (4)
Example: W = 10 lb.
A=Half lift=0-375".
n=400 R.P.M.
From (3) p=0-157 x 400 = 62-9.
From (4) X=;^ X 0-375 x62-92 = 38-5 lb.
The valve would therefore require to be provided with a spring
capable of exerting a force of 48-5 lb. to deal with inertia and
dead weight only, apart altogether from gas pressure and
friction.
The use of the harmonic cam, to which the above figm-es
apply, has not become very general in Diesel Engine practice,
the tangent cam shown in Fig. 228, Chapter XIII, being more
commonly employed. Typical velocity and acceleration
curves for a tangent cam are shown in Fig. 198.
Effect of Levers and Push Rods, etc. — The simple case
considered above of a cam operating directly on the end of the
valve is seldom realised in practice, and a somewhat more
general case, illustrated diagrammatically in Fig. 199, will be
considered. We have now several different members, all
participating in the motion and acquiring momentum which
232
DIESEL ENGINE DESIGN
AIR AND EXHAUST SYSTEM
233
must be overcome by the spring. The problem is a simple
case of the general theory of a system of one degree of freedom
simplified by treating the " coefficient of inertia " as a constant
instead of a function of x. Consider any particle of the lever
AB situate at a distance " r " from the fulcrum E and having"
k
E
^-jtAwtts^
^b\
PK^-J_
J
_C_
^
h
"^
\j ^■
i
Fig. 199.
a weight " w " lb. If s denote the speed of this particle during
any small displacement of the system, then : —
s=x.- and s=x.-
a a
The effective force acting on the particle is therefore equal to
•X— and the reaction Xj at A due to all such particles of the
g a
lever is given by
^ 2w.r2=?.Wi^i'
-(5)
ga^ g a^
Where Wi is the weight of the lever and k^ is its radius of
gyration about E.
W k 2
The expression — i-|- is the inertia co-
g.a
efficient of the lever with respect to the co-ordinate x, and may
be denoted by A^. The total reaction X at A, due to the inertia
effects of the valve itself, aU the levers, push-rods, etc., is the
sum of all the reactions due to the individual members, and
therefore, -j^^^ (Ao+A^+A.+A,) (6)
Wn
Aq being= — where Wo=weight of valve.
234 DIESEL ENGINE DESIGN
and Aj is the inertia coefficient of the push-rod, and A» is
that of the link.
By similar reasoning to that given above for A i it is found
W,/b\2
that A2= — -I - ) where W2=weight of push-rod.
and A,=^ — -I -^ ) , where W3=weight of link, and k3=its
g \ ac /
radius of gyration about its axis.
It will be seen at once that equation (6) is similar to (1),
with inertia coefficient substituted for mass. The assumption
made is that the angles a ^ydo not deviate far from 90 degrees.
For ordinary practical purposes a deviation of 10 or 15 degrees
on either side involves a negligible error.
Example: Wo= 6 lb. a=10". kx=7"
Wi=151b. b = 12". ka=5"
W2= 81b. c= 7".
W3= 51b. x = 1600in./sec.2
X=x(Ao+Ai+A2+A3)
1600r„ , 1 ./ 7 Ws/12\' , /12X5Y-
= 118 lb.
For slow running engines the inertia does not usually amount
to more than two or three lb. per sq. in. of valve area. In most
cases the spring may be based on the inertia load plus about
6 lb. per sq. in. of valve area, to deal with the other effects
which enter into the question.
Strength and Deflection of Springs. — ^The usual formulae for
the safe load and the deflection of springs made of steel wire
of circular section are given below for handy reference.
'P=Safe load (maximum) in lb.
Tg f= Safe stress, usually about 60,000 lb.
P=0-2f — Where i gper sq. in.
^ I d=Diameter of wire in inches.
r =Mean radius of coils in inches.
S =Deflection in inches.
n= Number of turns (free).
G=Modulus of rigidity, usually
taken to be about 12,000,000
lb. per sq. in.
And
64.n.r^ P
"^ d* G Where
AIR AND EXHAUST SYSTEM
235
Exhaust Piping. — A common arrangement of unjacketed
cast-iron exhaust piping is shown in Fig. 200. The flexibility
of the system renders any special provision for expansion un-
necessary. The piping itself being out of reach need not be
lagged, and may be as light as casting considerations will allow.
Fig. 200.
The connecting pieces between the various covers and the
common discharge pipe may be made equal in bore to the
diameter of the exhaust valve. The bore of the collector pipe
joining the silencer may be proportioned with reference to the
nominal velocity of the exhaust gases as follows : —
Let Vp =Piston speed in feet per second.
Vj=Nominal speed of exhaust in feet per second.
B =Bore of cylinder in inches.
d= Bore of exhaust pipe in inches.
n=Number of cylinders.
T"™ v.=y.(2)-x5
In using this formula " n " should be put equal to 4 in all
cases where the number of cylinders is equal to or less than 4.
The reason for this is that the gases discharge intermittently,
and an engine of one, two or three cylinders requires approxi-
mately the same size of pipe as a four-cylinder engine of the
same size and speed. The value of V^ varies from about 70 in
236
DIESEL ENGINE DESIGN
small engines to 110 in large. The pipe leading from the
silencer to the atmosphere may be made about 25% larger in
the bore. A somewhat neater arrangement, involving an inter-
mediate collector under the floor, is shown in Fig. 201. The
individual exhaust pipes must now be water-cooled, but there
is no objection to short, uncooled sections in way of the flanges
of sufficient length to accommodate the bolts. If the pipes
are cast, the thickness of the outer walls need not exceed about
J" to f " with good foundry work. The jackets may also be of
welded steel tubes or sheets.
Fig. 201.
In marine installations the exhaust is sometimes used to
furnish a supply of hot water for heating and other purposes,
and this may be achieved by providing the exhaust collector
with nests or coils of tubes through which water is circulated.
As a rule the exhaust from a four stroke marine engine is not
sufficiently hot to necessitate the provision of a water-cooled
silencer apart from the arrangements which have just been
mentioned.
Silencers. — As a rule four stroke Land Diesel Engines are
supplied with a cast-iron silencer, having a capacity of about
six times the volume of one of the working cylinders, and a
common type is shown in F^. 202.
The result is not always satisfactory, and better iresults are
obtained by using a large underground brick or concrete
chamber having twenty or thirty times the volume of one
cylinder. Wrought iron silencers, unless water-jacketed or
buried underground, usually give out a ringing noise, unless
the gases on entry are made to difEuse through a trumpet
arrangement similar to that described above under suction
pipes.
AIR AND EXHAUST SYSTEM
237
Two Stroke Engines. Scavengers.— If the efficiency of the
scavenging process could, be definitely ascertained in every
case, it would be a simple matter to calculate a suitable
capacity for the scavenge pump. In those engines of which the
results on trial suggest that this process is nearly perfect the
scavenge pump appears to have a stroke volume capacity of
about 1 -4 times the aggregate stroke volume capacities of the
cylinders which it feeds. It hardly
appears safe, however, to assume that
this allowance will necessarily be suffi-
cient in any proposed design, and some
constructors have adopted a higher
figure, 1-8 to 2, in their first attempts
at two stroke design. The scavenge
air pressure is a factor which decides
itself when once the details of port
and valve openings have been fixed,
and check calculations must be made
on some such lines as those indicated
in Chapter III, to make sure that
reasonable limits of pressure will not
be exceeded. The pressures obtained
in practice vary from about 3 to 7 lb.
per sq. in.
The work done by the scavenger is
all lost work and must therefore be
kept at a low figure. On the other
hand, the super-pressure above that of
the atmosphere with which a controlled
scavenge engine starts the compression
stroke, is a valuable factor in increas-
ing the power of a given-sized cylinder.
If simplicity is the main consideration and uncontrolled port
scavenge is adopted, it becomes necessary to be content with
a very moderate M.I.P., as the super-pressure obtainable at
the beginning of compression appears to be very limited with
the arrangements hitherto adopted on such engines.
Construction of Scavengers. — It is not necessary to deal
exhaustively with the details of scavenge pumps as these differ
little from the corresponding parts of L.P. steam engines.
The scavenger, or scavengers, are preferably driven off cranks
provided for the purpose on the main shaft, as in Fig. 203. In
Fig. 202.
238
DIESEL ENGINE DESIGN
marine designs the link-drive shown in Fig. 204 has also been
used, but as usually carried out is open to the following
objections : —
(1) The arrangement does not lend itself readily to the closed
engine type of framework which appears to have every
advantage for Diesel Engine work.
(2) The side levers involve cantilever connections, which
lack rigidity.
(3) The heavy reversals of thrust are liable to cause knock-
ing and vibration, owing to small bearing surfaces and
poor lubrication.
Fig. 203.
Fig. 204.
Fig. 205.
A scheme suitable for land work but a trifle inaccessible for
marine purposes consists of a tandem arrangement of scavenger
and low pressure stage of blast air compressor, as in Fig. 205.
Valve Gear. — ^The earlier Diesel scavengers were fitted with
piston valves single or double ported, and this type of gear is
still retained in some des^s. For reversing such valves the
Stevenson link motion appears to be the best solution on hand
at present. In other designs automatic disc or plate valves are
being increasingly used, with a gain in simplicity and the
advantage in marine work that no reversing gear is necessary.
The liability of the valves to failure by fatigue seems to be their
chief drawback.
Scavenge Air Receivers. — From the scavenge pump the air
AIR AND EXHAUST SYSTEM 239
passes to a receiver in. communication with the cylinder covers
or scavenge air belts of the working cylinders. With the usual
arrangement, where one, or at most two, double-acting scavenge
pumps are used to supply a number of working cylinders, say
three to eight, it is advisable to make the capacity of the
receiver large compared with the stroke volume of one
scavenger, in order that the pressure in the receiver may
remain sensibly constant, and a suitable capacity may usually
be secured by making the diameter of the receiver about
1 to 1-3 times the cylinder bore. In any proposed case it is a
simple matter to construct a diagram showing on a time base
the rate at which air is being passed to the receiver by the
scavenger and carried away from it by the working cylinders.
The resultant effect in creating fluctuations in a receiver of any
proposed capacity is then easily calculated. The effect of too
small a receiver capacity is to give those cylinders which begin
compression at the instant of maximum scavenge pressure an
advantage over those less favourably timed. The effect is
readily studied in practice by means of light spring diagrams
taken from the receiver and the working cylinders respectively.
By way of example, if two cylinders start compression with
absolute pressures of 21 and 20 lb. per sq. in. respectively, then
the first wiU (other things beiag equal) have a maximum power
capacity 5% greater than the second, or if they are worked at
the same power the second cylinder will work with a mean
absolute charge temperature 5% greater than the first, which
is no small evil. In some two stroke designs one double-acting
scavenger is provided for each pair of working cylinders, and if
the deliveries are correctly timed with respect to the scavenge
periods the receiver capacity does not require to be very large.
Scavenge receivers are usually made of riveted or welded sheet
steel, having a thickness of about one per cent of the diameter.
A disastrous explosion at Nuremberg in 1912, traceable to the
ignition of lubricating oil vapour in the scavenge receiver of a
large two stroke engine, points to the desirability of providing
drain cocks and a relief valve or safety diaphragm of large area.
The area of the trunk communicating with each cylinder
should be well in excess (usually about double) of the maximum
aggregate area of the valves or ports which it supplies.
Scavenge Valves. — In different designs embodying the
principle of scavenging through the cover one, two, three and
four valves have been employed. In particular, if two valves
240
DIESEL ENGINE DESIGN
are used they may be identical with the air and exhaust valves
used in four stroke engines of the same size, with the piston
speeds at present customary. With any other number of
valves the area may be made equivalent. In view of the
Fig. 206.
relatively light duty which devolves upon them, scavenge
valves are usually made of somewhat simpler construction
than exhaust valves, as shown in Fig. 206. To prevent loss of
scavenge air past the spindle the latter is usually made a good
fit.
With controlled port scavenge, double-beat valves, piston
AIR AND EXHAUST SYSTEM
241
valves and Corliss type valves have all been used in different
designs, as shown diagrammaticaUy in Fig. 207. The use of
the valve is exclusively to regulate the instant at which air is
admitted to the cylinder, the point of cut-off being determined
by the piston covering the ports on the up stroke.
It therefore foUows that so long as the valve is full open at
the instant when the ports are covered, the point at which the
valve seats again, may be determined arbitrarily by other
considerations affecting the valve gear.
Exhaust System. — The chief evil to be guarded against in
the design of the exhaust system of a two stroke engine is the
interference of the exhaust rush of one cylinder with the
scavenge process of another. An obvious but inconvenient
J way of avoiding this trouble is to provide separate exhaust
Fig. 207.
pipes and silencers for each cylinder. Practically the same
effect can be achieved by providing common exhaust systems
for pairs of cylinders whose cranks are at 180°. This, however,
does not appear to be necessary, and satisfactory results with
an exhaust system common to all cylinders are obtained by
the use of arrangement similar to that shown in Fig. 201 for
four stroke engines.
The essential point is that the pipes which connect each
cylinder to the common collector should be of large diameter
(about three-quarters of the cylinder bore), and that the
collector itself should have a volume large in comparison with
the stroke volume of one cylinder. It does not follow, however,
that a free passage for the exhaust is a necessary condition for
efficiency, as the latter has sometimes been improved by the
insertion of a throttling diaphragm in the exhaust passage at
the point where the pipe joins the cylinder.
242
DIESEL ENGINE DESIGN
Silencers. — The sudden release by the uncovering of ports
of a pressure of 40 lb. per sq. in. and upwards produces a noise
which in the absence of a silencer can be heard some miles off.
A type of sUencer which renders the exhaust inaudible a few
yards away without imposing any back pressure, is shown in
Fig. 208. This type of silencer is most effective when water-
Diameter of Silencer
. 3-5 B
Length ,, ,,
. 5-6 B
Bore of Inlet Pipe .
. OSS B
EEt
Fig. 208.
jacketed. Approximate main dimensions are given in terms of
the bore of the engine cylinder on the assumption of a piston
speed of about 800 feet per minute. For land purposes a large
pit without special baffles would probably serve equally well.
Literature. — For information on the mechanics of cam-
operated mechanism, see : —
Goodman, J., Mechanics Applied to Engineering (Longmans).
CHAPTER XII
COMPRESSED AIR SYSTEM
As mentioned in Chapter I, the injection of fuel by means of
an air blast is one of the outstanding characteristics of the
Diesel Engine, and it seems probable that its use in the early
experimental engines was suggested by the compressed air
apparatus used to start the engine, and which still appears to
be the most practicable method of doing this. The utihty of
the air blast is by no means confined to its function of injecting
the fuel ; in fact the widespread use of mechanical means of
injection in other types of oil engine clearly indicates that
effective atomisation can be obtained otherwise. As Guldner
has pointed out, the use of an air-blast probably secures a
more efficient mixing of the cyhnder contents than could be
obtained in any other practicable manner. The advantages in
efficiency which such mixing secures are easily appreciated on
examination of the thermodynamic principles involved. With
good mixing the combustion proceeds rapidly, and reduces
after-burning to a minimum ; and, further, the whole charge
tends to remain homogeneous as to temperature, a necessary
condition for maximum efficiency. In practice there are two
aspects from which the efficient utilisation of heat should be
viewed, viz. : —
(1) Economy in fuel consumption.
(2) The effect of efficient combustion in keeping the mean
cycle temperature to a minimum.
The last consideration is a vital one from the point of view of
rehabUity and durability, and experience abundantly proves
that a relatively small increase of the cycle temperature, due
to overloading, leakage past valves, loss of volumetric effi-
ciency or other causes is sufficient to convert an otherwise
reliable machine into a source of continual trouble.
243
244 DIESEL ENGINE DESIGN
The blast air has a pressure which varies from about 900 to
1000 lb. /sq. in. at full load, to about 600 lb. /sq. in. at no load,
and in land engines is usually supplied by a comprescor form-
ing an integral part of the engine. In marine installations the
compressors are sometimes driven by separate auxiliary engines.
The arrangement adopted by one maker is to drive the lower
stages of the compressors by auxiliary engines, the last stage
being performed by a high pressure plunger driven by the
main engine. From the compressor the air passes, via coolers,
to a blast reservoir or bottle of sufficient capacity to absorb
fluctuations of pressure, and fitted with suitable distributing
valves, one of which communicates with the fuel injection
valves and another enables surplus air to be passed to the
storage reservoirs provided for starting purposes. Cam-
operated valves in the covers of one or more cylinders enable
the stored air to be used, to give the engine the initial impetus
which is necessary before firing can begin. With land engines
the starting bottles are generally charged to a pressure of
about 900 lb. per sq. in., and with the fly-wheels commonly
used it is not necessary to provide starting valves for more
than one cylinder out of three. With marine engines, storage
pressures of about 300 lb. per sq. in are more common, on
account of the difficulty of making high pressure reservoirs of
large size, and in order to secure prompt starting from any
position starting valves are fitted to every cylinder.
The air system also includes certain servo-motors or air
engines,, frequently used to perform operations of reversing
the valve geiir. The various organs will now be considered in
more detail.
Air Compressors. — Four stroke land engines of the slow
speed type as at present constructed require compressors
having a capacity of about 15 cubic feet per B.H.P. per hour,
which assuming a volumetric efficiency of 80% corresponds to
a stroke volume capacity|rf)f about 19 cubic feet per hour.
High speed engines appear as a rule to require about 25% more
than this allowance. The above method of basing the com-
pressor capacity on the B.H.P. is not a very satisfactory one,
as different makers have different views as to power rating.
A better plan is to express the L.P. stroke volume as a per-
centaige of the aggregate cylinder volume, and the following
figures are representative of average practice for land engines.
In view of the demands made on the system when manoeuvring,
COMPRESSED AIR SYSTEM
245
marine engines ai;e usually provided more liberally, to the
extent of 50 to 100%.
Bore of
Ratio L.P. stroke vol. -H Stroke vol. of working
cylinders.
working cylinders.
Four Stroke Engines.
Two Stroke Engines.
10
15
20
0-08
0-07
0-05
0-16
0-14
0-09
Number of Stages. — ^For small slow-running compressors
two stages are sufficient, but a 9-inch diameter of low pressure
■cylinder appears to be about the safe limit ; and even with
this restriction it appears wise to abandon the principle of
equal distribution of work between the stages. The small
diameter of the H.P. cylinder affords little cooling surface for
the dissipation of heat, and this consideration points to the
advisability of arranging for the greater part of the work to be
done in the L.P. stage, a conclusion which has been anticipated
by experience.
Three-stage compressors are being increasingly used, even
the smaller sizes, four stages being required in the very largest
installations only. With three or four stages the principle of
equal division of work is open to less objection, owing to the
smaller ratio of compression in each stage.
Compressor Drives. — ^Almost every conceivable type of
drive has been adopted at one time or another, and only the
commonest are mentioned below : —
(1) Tandem two or three stage compressor driven off the
crank-shaft. This arrangement appears to have the
balance of advantages for most purposes.
(2) Tandem two-stage compressor driven by links and levers
from each connecting rod or crosshead. This arrange-
ment is expensive, but has the advantage of distribut-
ing the work amongst a number of small compressors,
which are subject to less heat trouble than one com-
pressor of the same capacity. The suction pressure of
the H.P. stage is usually sufficient to prevent reversal
of thrust due to inertia, and consequently sweet run-
ning is secured.
246 DIESEL ENGINE DESIGN
(3) Similar to (2), but stages separate. This arrangement is
bad, as tlie cooling siirface is less than in case (2), and
the L.P. gear is subject to reversal of thrust due to
inertia.
(4) Twin tandem cylinders driven by links off the crank-
shaft. This arrangement gives good results, as the
load is divided between two units and the pressure on
the crank-pin is reduced by the leverage of the link-
work.
Constructive Details. — ^The construction of air compressors
being a specialised branch of mechanical engineering, it is not
proposed to give here more than a very brief reference to the
subject . The cylinders of tandem two and three stage machines
are frequently cast in one piece, including the water-jacket.
The relatively low temperatures obtaining justify this pro-
cedure, provided sound castings can be obtained with reason-
able regularity. The foundry work may be simplified in the
case of two-stage compressors by the following division of
material :—
L.P. Cylinder and Jacket — one casting.
L.P. Cover and H.P. Jacket — one casting.
H.P. Liner and H.P. Cover — separate castings.
One advantage of this scheme is the possibility of renewing the
H.P. Liner when worn. The latter is peculiarly subject to rapid
wear on account of the high pressure behind the rings.
Similar arrangements are of course possible with three-stage
machines. The possibility of increasing the cooling surface of
the H.P. liner by means of ribs does not appear to have received
very much attention. The trunk pistons serve as admirable
crossheads, being almost entirely free from the heat troubles
to which the pistons of internal combustion
engines are subject. For the L.P. and inter-
mediate stages Ramsbottom rings are usually
fitted. In j^ery large machines the latter can
also be fitted to the H.P. plunger. Small H.P.
plungers are usually fitted with some arrange-
ment similar to that shown in Fig. 209. The
only thing which need be said about the
connecting rods is that on account of the
thrust being always in one direction, special
care is required in the details of lubrication.
The design of valves has an important bearing on the success
COMPRESSED AIR SYSTEM 247
or failure of a compressor. The chief evils to be avoided are : —
(1) Sticking of the valves off their seats, due to deposits of
carbonised oil.
(2) Damage to valves or valve seats, due to hammering.
The first is influenced more by the efficiency of the cooling
arrangements and the compression ratio than with the design
of the valves themselves. For obvious reasons the H.P. valves
are most subject to this trouble.
The second trouble is usually due to the valves being too
heavy, having too much lift, or the failure to provide adequate
cushioning, and in successful designs is avoided by one or more
of the following means : —
(1) Making the valves in the form of very light plates or discs -
with a very small lift.
(2) Providing a large number of very small valves in place
of one or two large ones.
(3) Where large valves of considerable weight are used,
arranging for some sort of dash-pot action.
All the valves should be easy of access and removal. Ex-
periments with existing types of compressor seem to indicate
that makers are inclined to base their valve dimensions on an
air speed very much lower than is necessary. One or two per
cent loss of efficiency is of small importance, if such a sacrifice
enables the size of the valves to be reduced.
In some designs the intercoolers are separate from the
compressor cylinder, and in others take the form of pipe-coil^
arranged round the compressor cylinders inside a removable
water-jacket. Vibration of the coUs should be prevented by
adequate clamps and stays, and no sharp bends are allowable,
on account of a scouring action (presumably due to turbulent
flow) which in acute cases may cause fracture of the pipe in a
short time. L.P. intercoolers are sometimes made similar to
tubular condensers, and in other designs take the form of a
cast-iron vessel provided with internal helical baffles which
give rise to turbulent flow and increase greatly the efficiency of
the cooling surface. It is desirable in all cases to fit a final
cooler, to reduce the temperature of the fully compressed air
before entering the blast receiver. Each receiver should be
fitted with safety valve and drain. One or two isolated cases
of explosion, traceable to accumulation of lubricating oil in the
intercooler system, emphasise the necessity for these fittings.
248 DIESEL ENGINE DESIGN
Some makers fit special " purge-pots " in communication with
each receiver for the collection and discharge of condensed
water and oil.
Calculations for Compressors. — In calculating the L.P.
stroke volume required to furnish a given free air capacity,
allowance must be made for the volumetric efficiency, which
depends mainly on the clearance space and the delivery
pressure. For example, suppose the clearance to be 3% of the
stroke volume and the receiver pressure to be 150 lb. per sq. in.
On the suction stroke, the suction valve will not begin to lift
until the air left in the clearance space has expanded down to ,
atmospheric pressure. If this clearance air expands according
to the law : —
P.V.i-2=constant,
then its expanded volume expressed as a percentage of the
stroke volume wiU be : —
/I HA .n\ 1
= 22-5%
Subtracting its original volume, viz., 3%, the amount by which
the effective stroke is shortened is 19-5%, and the volumetric
efficiency is 80-5%. A further deduction should strictly be
made for the fact that at the beginning of the compression
stroke the cylinder contents are in general at a pressure
slightly less than atmospheric. One or two per cent will usually
cover this contingency. This example is sufficient to show the
importance of reducing the clearance volume of the L.P.
cylinder to a minimum. The efficiencies of the L.P. or H.P.
cylinders may be found similarly, but only influence the
volumetric efficiency of the compressor as a whole indirectly
by raising the receiver pressure above the value it would have
if there were no clearance. It will be evident on reflection that
leakage past the H.P. delivery valves will also raise the receiver
pressures, and for this reason it is desirable to fit pressure
gauges to all receivers so thsft the condition of the valves may
be inferred from the gauge readings.
Assuming perfect intercooling and equal volumetric effi-
ciency in all the stages, the pressure of the atmosphere and the
receiver pressures (absolute) will be in inverse ratio to the
stroke volumes of the cylinders, and equal division of work
between the stages will be secured by the proportions given
below ;■ —
COMPRESSED AIR SYSTEM 249
Two f Atmospheric Pressure. Intermediate Pressure. High Pressure.
.\ 1 at. -14-7 lb./in.2 8 at. =118 Ib./in.^ 64 at. -940 Ih./in.'
T
L.P. volume. -8 H.P. volume.
Atmospheric 1st Intermediate 2nd Intermediate High
Pressure. Pressure. Pressure. Pressure,
lat. =14-71b./ 4at. -58-81b./ 16 at. =235 lb./ 64at. =940 lb./
in.* in.= in.^ in.^
Three
Stages.'
L.P. volume : I.P. volume : H.P. volume =16 : 4 : 1.
In practice better results are obtained with two-stage
machines by the following proportions : —
Atmospheric Pressure. Intermediate Pressure. High Pressure.
1 at. =14-7 lb./in.» 12 at. =177 Ib./in.^ 64 at. -940 Ib./in.^
L.P. volume -12 H.P. volume.
Assuming that the L.P. stroke volume has been determined by
some such considerations as the above, the actual cylinder
dimensions are found by selecting suitable values for the
piston speed and the L.P. stroke to bore ratio. The piston
speeds commonly used lie between about 300 to 600 feet per
minute, the lower speeds being usually associated with small
machines. With a little care it is possible, by making small
variations in the strokes, to design a series of four or five
compressors suitable for engines covering a wide range of
powers. Such a scheme involves sacrifices in some cases which,
however, would appear to be quite outweighed by the advan-
tages of standardisation. The ratio of stroke to bore of the
L.P. cylinder wiU generally lie between about 0-7 and 1-7.
The valve areas for each stage are based on some figure for the
mean velocity obtained, in the case of suction valve, by multi-
plying the mean piston speed by the ratio of piston to valve
area, and in the case of delivery valves the mean piston speed
during the delivery period by the same ratio. Certain con-
tinental authorities recommend speeds not exceeding 80 and
115 feet per second for the suction and delivery respectively,
but it appears that these figures may be doubled or even trebled
with impunity, and sometimes to advantage.
The calculations of the strength of the various parts are
straightforward, involving no special principles, and are there-
fore passed over. The cooling surface to be provided for inter-
cooling is a very important matter, and the following figures
from a eucoessf ul design may be useful as a basis of comparison
in the absence of first-hand experimental data.
250
DIESEL ENGINE DESIGN
Three-stage Compressor,
Cooling surface,
copper-pipe coils.
Free air capacity, 130 ft. ^ /min.
L.P. 8-7 ft. 2
L.P. 3-6ft.2
H.P. 3-6ft.2
Fig. 210.
The subject of heat transmission being a very important one
■^ ^ in connection with internal combustion
engines, a brief reference to the usual
theory is inserted below.
Transmission of Heat through
■ Plates. — Referring to Fig. 210, the
r^ direction of heat flow is indicated by
;. ■ the arrow and the symbols t,^, t^, tg, t^
denote the temperatures of the hot
fluid, the hot side of the plate, the cold
side of the plate, and the cold fluid
respectively. The total heat drop con-
sists of three stages : —
(1) An apparently sudden drop from the hot fluid to the
plate.
(2) A steady gradient across the thickness of the plate.
(3) An apparently sudden drop from the plate to the cold
fluid.
The assumption is that the rate of heat flow is dependent
only on the temperature drop, the thickness of the plate, and
the particular fluids employed. On this assumption, if Q is
the amount of heat transmitted per hour per unit of area,
then : —
Q=ai(ti— t2)=^(t2-
t3)=a2(t3-t4) (1)
from which
Q=
(ti-t*)
-(2)
1 +i-F-
Where d=thickness of plate and oi, 02, and X are constants.
According to Hiitte, for air at atmospheric pressure and
small velocities ai=about 0-4:-{-1-1-\/y.
Where v= velocity in ft. /sec, and for water (not boiling
and without turbulence)
a2=about 100 B.T.U. /ft.^ deg. F.
COMPRESSED AIR SYSTEM 251
Values of X are given below for various metals : —
Iron . . . 460]
Mild steel . . 320 1 B.T.U. /ft. ^ deg. F. per inch of
C!opper . . . 2100 j thickness.
Brass . . .740/
The values of X are well determined, but unfortunately
the term involving this constant is the least important of the
three as the bulk of the heat drop occurs at the surfaces of the
plates. The values of oi and a 2 must be used with caution as
the small amount of published data indicates that these
constants are subject to enormous variation under different
circumstances. In particular, the value of a 2 is greatly
increased by the eddying motion produced by the introduction
of spiral baffles in condenser tubes and the like, and is increased
about tenfold if the water boils. The value of a^ for air
increases greatly with rise of temperature and pressure.
Apart from these uncertainties, formula (2) suggests the
foUoTving corollaries : —
(1) The rate of heat transmission is but little influenced by
the thickness of the plate in most practical cases,
such as cylinder covers, etc.
(2) The temperature drop across the plate, and consequently
the temperature stress, is proportional to the rate of
heat transmission and to the thickness of the plate ;
this suggests one reason why it is advisable to work
large engines at a smaller mean cycle temperature
than small ones.
For further information on the subject of heat transmission,
the reader is referred to the sources of information mentioned
in the footnote.^
Air Reservoirs. — The usual arrangement of air reservoirs for
land engines is shown in Fig. 211. This scheme was devised
in the very early days of the development of the Diesel Engine
and no substantial improvement has been made on it in recent
years. Two starting and one blast-air bottles are provided,
all designed for a working pressure of about lOOO'lb. per sq. in.
One of the starting bottles serves as a reserve, in case of a
1 " High-speed Internal Combustion Engines," Judge. " Heat Trans-
mission," Report by Prof. Dalby to the Inst. Meoh. Engs., 1909. "Notes on
Recent Researches," paper by Prof. Petavel : Manchester Assoc, of Engineers,
Oct., 1915. "The Laws of Heat Transmission," Lecture by Prof. Nicholson :
Junior Inst., Jan., 1909.
252
DIESEL ENGINE DESIGN
failure to start the engine, due to any derangement. In the
event of such failure, every care is taken to make certain that
the engine is in perfect order before using the reserve bottle,
and it seldom happens in practice that the bottles require to
be replenished from outside sources of supply. The connections
between the bottles, the air compressor and the engine should
Starting Pipe
jmpressor-
To Fuel Injection Valves
Fig. 211.
be quite clear from the diagram. Only one or two points
will be mentioned.
Before starting up, it i^ossible to ascertain the pressure
in each of the three bottles by opening up the appro-
priate valve on each bottle-head in rotation In each
case the pressure is recorded on the left-hand gauge..
The pressure in any pair or all three bottles may be
equalised by opening up a pair or all three such valves.
The right-hand gauge registers the blast pressure on the
engine side. By throttling the blast control valve on
(1)
(2)
(3)
COMPRESSED AIR SYSTEM
253
the blast bottle-head the injection pressure may be
regulated below that of the bottle. This is done when
replenishing the starting bottle on light load. It is
thus possible to pump up the starting vessel to
1000 lb. /in. 2 whilst the blast pressure is only 600
lb. /in. 2, as required for light running.
FlQ. 212.
The bottle-heads containing the various valves are usually
machined from a solid block of steel A detail of one of the
valves is shown in Fig. 212.
The bottles themselves are of weldless steel, and a neck is
frequently screwed on, as in Fig. 213. Some idea of the
capacities of the bottles commonly provided may be gathered
from the following table : —
254
DIESEL ENGINE DESIGN
Total Capacity of H P Starting Air Bottles
(four stroke engines)
Engines of about 9" bore, having one to six cylinders — about
fourteen times the stroke volume of one cylinder.
Engines of about 24" bore, having one to six cylinders — about
seven times the stroke volume of one cylinder.
Owing to the expensive machinery required to manufacture
weldless reservoirs, only a certain limited number of standard
sizes are available at reasonable prices, and in very large instal-
lations it is sometimes necessary to provide groups of four or
more starting vessels. The usual working stress is about
Fig. 213.
8000 lb. /in.^, and it is customary to specify a water-test
pressure of double the working pressure.
Riveted Air Reservoirs. — For marine installations of high
power it is usual to use a lower air pressure of about 300 lb. /in.^
for starting purposes. The air reservoirs now require to have
a very much larger cubic capacity, but the reduced pressure
permits of the employment of riveted reservoirs. The con-
struction of the latter need not be dealt with here, being com-
parable with that of the steam drums of modern water-tube
boilers. Adequate drainaps for condensed water and oil
vapour, and also a manhole for inspection and cleaning, should
be provided. These matters, as well as others dealing with the
strength of the riveted joints, the quality of material to be
used and the tests to be carried out on completion, form the
subject-matter of regulations by the various insurance societies
and the Board of Trade.
Blast Piping System. — From the blast bottle the injection
COMPRESSED AIR SYSTEM
255
air passes to a main running along the back of the engine,
where it is distributed by short lengths of pipe to the several
fuel valves, as in Fig. 214. Where one fuel pump is provided
for a number of cylinders the fuel distributors may be made to
serve as distributing tee-pieces for the blast air. In marine
^
Fig. 214.
engines it is usual to provide a shut-down valve, as in Fig. 215,
to each tee-piece, so that the supply to any individual cyhnder
may be cut off, to enable the fuel valves to be changed without
stopping the engine.
In addition, it is sometimes necessary (see Chapter XIII)
to provide a valve whereby the whole supply of blast air
is automatically cut off when the
manoeuvring gear is put into the
stop position. The bore of the blast
air supply pipe' to each cylinder
need not exceed about 2% of the
cylinder bore, but is usually greater
than this in small engines, to avoid
the multiplication of standard sizes
of unions. The blast air main may
be about 4% of the cylinder bore
for any number of cylinders up to
about six. The same type of union
may be used as has already been -
illustrated in Fig. 156 in connection
with the fuel system. Other types
of union are in use, notably the
Admiralty Cone Union, which is also
very serviceable*. Fiq. 215.
Blast Pipe
256
DIESEL ENGINE DESIGN
The Starting Air Pipe System. — With land engines it is
quite common to provide one cylinder only with a starting
valve when the number of working cylinders does not exceed
miiiuuk
, from ^/r Bottles
Fig. 216.
\l.Vl.\^VW
Fig. 217.
four. With six cylinders and upwards two and sometimes three
units are provided with air-starting arrangements. A neat
arrangement of the starting pipe is shown in Fig. 216 for a
four-cylinder engine. In this case the design of the starting
valve is such that air is admitted through a port cast in the
side of the cylinder cover. Any arrangement of piping is to
be avoided which renders difficult the removal of a cylinder
cover, hence the provision of an elbow on the latter. In other
designs this elbow is cast integrally with the cover itself.
COMPRESSED AIR SYSTEM
257
With marine engines all cylinders are provided with starting
valves, to which the air is led through a steel main pipe line
running the whole length of the engine. Fig. 217 shows the
type of pipe flange most commonly used, the material being
steel. The tee-pieces for distribution to the several cylinders
may be of cast iron or cast steel. If the former material is used,
the design should be very substantial, as in Fig. 218. The pipe
lines should be securely clipped to the framework of the engine,
otherwise there is liable to be severe vibration, due to the
surging of pressure within the pipe.
In large slow speed engines the diameter of the starting pipe
may be about 0-07 to 0-1 of the cylinder diameter. In small
high speed engines it is advisable to give the main distributing
pipe a diameter of about 0-15 to 0-17 of the bore, in order to
secure rapid acceleration.
Starting Valves. — These are usually
located in the cylinder cover and
operated by cams and levers, in the
same way as the other valves. An
arrangement less frequently used con-
sists of a centralised air distributing
box of rotary or other type remote
from the cyUnder covers but connected
to them by distributing pipes. Loss of
compression is obviated by the pro-
vision of non-return valves in the ■
cylinder cover. The centralised dis-
tributing box may consist of a sleeve
rotating in a casing in such a manner
that a slot in the sleeve admits air
successively to a number of ports com-
municating with the several cylinders.
In other arrangements a set of cam-
operated mushroom valves is used.
These schemes have not become com-
mon practice and will not be discussed
here in further detail.. ^'''- ^l^-
A common type of starting valve is shown in Fig. 219. No
provision has been made here to prevent leakage past the
spindle, and if the latter is a good ground fit in the casing the
leakage should not be serious in amount. In large sizes of
valve additional tightness may be secured to advantage by
258 DIESEL ENGINE DESIGN
the provision of a number of small Ramsbottom rings. It is
usual to make the diameter of the piston part of the spindle
the same as the smaller diameter of the valve head. The
minimum spring compression should be equivalent to the
maximum starting air pressure acting on an even area equal
to that of the valve seat. The valve casing should be sub-
stantially proportioned, to prevent distortion and consequent
leakage at the seat or binding of the spindle. It will be noticed
that with this design of valve, joints have to be made at A and
B simultaneously. There is no practical difficulty about this.
Joint A is usually made with a copper or white-metal ring.
A slight modification is sometimes made by the introduction
of two cone joints, as in Fig. 220. This also works well.
Fig. 221 shows a type of starting valve in which the air is led
to the top of the valve casing, instead of being introduced
through a port in the cylinder cover.
A useful type of starting valve, devised by the Burmeister
& Wain Company for Diesel Marine Engines, is illustrated
diagrammatically in Fig. 222. With this design the valve
becomes inoperative when the air pressure is removed and
resumes working as soon as the pressure is restored. This
results in a great simplification of the manoeuvring gear (see
Chapter XIII) by the elimination of mechanism which in some
other designs is provided for the purpose of throwing the
starting valves out of gear when the fuel is turned on. The
desired result is achieved by attaching to the upper end of the
valve spindle an air cylinder and piston, kept in constant
communication with the air supply by means of holes through
the spindle. In the absence of air pressmre, the spring A is
sufficiently strong to keep the piston B at the bottom of the
cylinder C, thus removing the roUer from the range of opera-
tion of the cam D. When pressure air is turned on the piston
is forced to the top of the cylinder, and the valve remaias
operative so long as the force required to open the valve is less
than the difference betweeft the pressure load and the spring
load on the piston B.
Diameter of Starting Valves. — On theoretical grounds,
the necessary diameter of startiag valves would appear to
depend on the pressure of the air supply, amongst other things.
It so happens, however, that in those cases where a low pressure
air system is the most convenient (viz. in large marine instal-
lations) the multiplicity of cylinders to which starting air is
COMPRESSED AIR SYSTEM
259
II
w
§•
Fig. 220.
Fig. 221.
Fig. 222.
260
DIESEL ENGINE DESIGN
supplied, affords adequate starting torque with a relatively
low mean starting pressure in each cylinder. The result is
that roughly the same diameters of starting valve are used in
either case, i.e. whether a low or high pressure starting system
be adopted, typical figures being from about 0-1 of the bore
in the case of large engines to 0-13 in small engines.
With slow speed land engines it is very desirable to obtain a
" fat " starting card, to overcome the inertia of the heavy fly-
wheels which are usually necessary. An engine in good work-
ing order should start firing in the first or second revolution on
starting up cold. It seems probable that in the event of low
pressure air being used for such engines, it might be necessary
to fit starting valves to aU the cylinders or to make the dia-
meter of the latter larger than is customary with the high
pressure starting air systems at present in use.
M
Oil
Cyjinder
Cocks for changing over From
Pneumatic to Hand Operation
Hand Pump
Oil Tank
Air
Cylinder
From Compressed Air Supply
^m. 223.
Air Motors. — In large marine engines the work required to
effect reversal of the valve mechanism when going from ahead
to astern, or vice versa, is generally too great to be done with
sufficient rapidity by a hand gear, except in case of a break-
down of the air motor which is usually provided for the
purpose . In different designs the air motor takes various forms,
of which some are mentioned below : —
COMPRESSED AIR SYSTEM
261
(1)
(2)
A small reciprocating engine (double acting), with two
cylinders and cranks arranged at right angles. The
arrangement is almost exactly similar to the small
auxiliary steam engines used on steamships for the
reversing gear or the steering gear. Low pressure air
is used, and the air motor is geared down by worm and
worm-wheel, so that it makes a considerable number
of revolutions for one movement of the reversing gear.
A single cylinder, with piston and rod, the reversiag
motion being performed in one stroke. This arrange-
ment is suitable for high or low pressure air, and in
either case the piston-rod must be extended into an oil
dashpot cylinder to reduce shock. If the two ends of
Fig. 224.
the oil cylinder be connected to a hand pump, the
latter may be used for reversing, in the event of a
failure of the air cylinder. This scheme is illustrated
diagrammatically in Fig. 223.
The reciprocating motion of the piston-rod may be
converted into rotary motion (one complete revolution
or more) by a rack and pinion.
(3) A rotary engine of the type which is frequently used as
a pump in connection with machine tools, motor-cars
and other small machines, and which is shown dia-
grammatically in Fig. 224. This type of motor is only
suitable for low pressures and is arranged to do its
work in a considerable number of revolutions by
means of worm gearing.
Types (1) and (3) would appear to have the advantage of
greater adaptability to varying pressures. It is an easy matter
262 DIESEL ENGINE DESIGN
to gear the motor down so that it will turn under the lowest
air pressures anticipated. At higher pressures wire drawing
at the ports prevents the attainment of an undesirably high
speed.
In type (2) hydraulic leather packings are used, and consider-
able care should be taken in the design and the workmanship
to eliminate aU unnecessary sources of friction. The strict
alignment of the two cylinders and the guided end of the rod
deserve special attention.
Literature. — ^Ford, J. M., " High Pressure Air Compressors."
— Paper read before the Greenock Assoc, of Shipbuilders and
Engineers. See Engineering, October 20th, 1916, et seq.
CHAPTER XIII
VALVE GEAR
Gams. — ^With few exceptions the valves are operated by
external profile cams made of cast iron chilled and ground on
the face. In order to facilitate the removal of valves and
cylinder covers, it is usual to arrange the cam-shaft to one side
of the cylinders and to transmit motion from the cams to the
valves through levers or a combination of levers and push-rods
or links. The arrangements in general use give an approximate
one to one leverage between cam and valve, and the figures
I Intone ■ OSS 0.
Fis. 225.
given above for the width of cam face are based on this pro-
portion.
Two forms of cam body for the air and exhaust valves of
four stroke non-reversible engines are illustrated in Fig. 225,
the dimensions being expressed in terms of the cylinder bore.
Fig. 226 shows a combined ahead and astern cam for a large
marine engine. The bosses should be bored a hard-driving fit
on the cam-shaft, and their lengths should be machined
accurately to dimensions, so that the complete group of cams
required for one cylinder give correct spacing when driven
hard up side by side.
The fuel cam has to be of special construction, on account of
263
264 DIESEL ENGINE DESIGN
the necessity for precise adjustment of the timing, and a
typical form is shown in Fig. 227. The toe-piece is preferably
made of hardened steel, but chilled cast iron is sometimes
used.
Profile of Cams. — ^The design of cam profiles for air, exhaust
and scavenge values is a matter of reconciling the claims of
the following desiderata : —
(1) Rapid and sustained opening.
(2) Absence of wear and noise.
In slow speed engines the question of noise hardly arises,
and wear is easily kept to a reasonable minimum by adequate
width of face. For such engines the tangent cam shown in
Fig. 228 is suitable. For high speeds a smoother shape, as
shown in Fig. 229, is desirable, and the relatively slow opening
may be compensated by earlier timing. Such profiles are
easily drawn by deciding on some arbitrary smooth curve of
roller hft, and J)lotting corresponding positions of the roller
with respect to the cam, as in Fig. 230. Some designers are in
favour of a sinusoidal form of roller-lift curve. With these
smooth profiles peripheral cam speeds of five feet or more per
second can be used with quite sweet running. On theoretical
grounds the cam profile should be based on the roller clearance
circle, as in Fig. 231, but it does not yet appear quite clear
whether the procedure has the practical advantages claimed
for it.
The starting air cams are best given a sudden rise on the
opening side to minimise wire-drawing (Fig. 232).
Combustion valve cam profiles are a study in themselves,
and the final decision rests with the test-bed engineers. The
effective period is usually about 48 or 50 crank-shaft degrees,
or 24 to 25 cam-shaft degrees. The cam-piece should, how-
ever, give a range about 25% in excess of this, after allowing
for the normal roller cleara^nce to allow for lost motion in
the gear. The tangent prorae shown in Fig. 227 is usually
found quite satisfactory, but other shapes are used.
In order to avoid noise in two stroke engines, it appears
necessary either : —
(1) To make the cams of smaller diameter than those of a
four stroke engine of the same size, or
(2) To provide double-faced cams mounted on a half-speed
cam-shaft.
VALVE GEAR
265
Fig. 230.
Fig. 231.
266
DIESEL ENGINE DESIGN
The first procedure is the more usual, but the second would
appear to have much in its favour. In one marine design
advantage is taken of this arrangement to work the engine on
the four stroke cycle at slow speeds.
Cam Rollers. — Engines having longitudinally fixed cam-
shafts are usually provided with cam rollers of steel case-
FlG. 232.
Fio. 233.
hardened and ground inside and out (Fig. 233) and having a
diameter of about one-third that of the corresponding cams.
The grooves provided for hand lubrication of the pin should
be noted. In marine engines in which reversal of rotation is
effected by the provision of ahead and astern cams mounted
on a longitudinally movable shaft, the roUers require to be
of large diameter (about 60% of the cam
diameter), in order that the idle cam may
clear the lever, as in Fig. 234. RoUers of this
size may be of cast iron bushed with phos-
phor bronze.
Valve Levers. — A common arrangement of
valve levers and lever fulcrum shaft for four
stroke land-engines is shown in Fig. 235.
The fulcrum brackets are secured to the
cylinder cover, and the latter may be lifted
complete, with all valves and gear, and re-
placed without disturbing the valve settings.
With the arrajigement shown it is necessary
to lift away the fulcrum shaft and levers
before the various valves can be removed
Fig. 234. for regrinding. This very slight inconvenience
VALVE GEAR
267
is sometimes overcome by means of split levers or by provision
of horse-shoe shaped distance collars on the fulcrum shaft,
which when removed leave sufficient space to aUow the levers
to be moved sideways clear of the valve casings. These devices
are desirable in the largest engines only. Referring to Fig. 235
below, it will be noted that the fuel and starting levers are
mounted on an eccentric bush A, connected to the handle B.
The latter is provided with a spring catch engaging with
notches in the fixed disc C, in accordance with the following
scheme and the diagram shown in Fig. 236.
Top notch. — Running. — Fuel lever in its normal running
position. Starting valve roller
out of range of cam.
Middle notch. — Neutral. — Both fuel and starting valve
rollers out of range of cams.
Fulcrum Spindle
For Valve Levers "
Fig. 235.
268
DIESEL ENGINE DESIGN
Fuel Needle
Valve N>
Starting 1
Starting
Position
Neutral
Position
(Running
Position
Fig. 236.
VALVE GEAR
269
Bottom notch. — Starting.-
-Starting air valve lever in its
working position. Fuel valve
roller out of range of cam.
Sometimes this arrangement is modified, by kejdng the
eccentric bush and handle to the fulcrum shaft and allowing
the latter to turn in its supports. This scheme is useful when
the disposition of the gear is such
that an air suction or exhaust
lever separates the fuel and start-
ing levers. This eccentric mount-
ing of levers may also be used '
for exhaust lifting or to remove
all the levers out of range of the
cams during the axial displace-
ment of the cam - shaft of a
reversing engine.
Certain well-known marine makers of great repute do not
consider it necessary to put the fuel valve out of action whilst
the engine is running on compressed air, and content them-
selves with suspending the supply of fuel to the valves during
this period.
The levers are generally of cast steel or malleable iron, but
good cast iron may be used if the stress is confined to about
Fio.
Fig. 238.
2000 lb./in.2 Some alternative sections are shown in Fig. 237.
The forked end of the lever calls for very httle comment.
Type A (Pig. 238) is a good design, but expensive. Type B is
very commonly fitted and is open to little objection. Type C
is the cheapest existing construction, and if accurate castings
(machine moulded) are obtainable the only machining opera-
270
DIESEL ENGINE DESIGN
tion required is to driU and reamer the hole for the roller-pin.
The two grooves for the taper -pin may be cast.
In small engines the tappet-end may consist of a plain boss
screwed to receive a hardened tappet-screw and lock-nut,
as in Fig. 239. In larger engines the more elaborate arrange-
ments shown in Fig. 240 are usually adopted. The bosses of
the levers should be bushed with good phosphor bronze and
provided with a dustproof oil cup of some description, in order
to reduce wear to a minimum. With these precautions the
bush should only require renewal at widely distant intervals
Glass Hard
Fig. 239,
Fig. 240.
and means of adjustment are unnecessary even in the largest
sizes of engines.
Strength of Valve Levers. — In four stroke engines the
exhaust- valve lever is the most heavily loaded. Although the
force required to operate the suction valve is relatively small,
it is usual to make the inlet valve lever of the same section as
the exhaust lever for the sake of uniformity of appearance, and
the same pattern may frequently be used for both. The loads
imposed on the tappet-ends of the various levers at the points
of valve-opening are given bdlow : —
FoTTE Stroke Engines
Exhaust Valve. About 45 lb. per sq. in. of exhaust valve
area+spring load+inertia of valve.
Suction Valve. Spring load +inertia of valve -j-a maximum of
about 5 lb. per sq. in. of valve area if the
exhaust valve happens to be closing too
early, due to excessive roller clearance.
VALVE GEAR
271
Starting Valve. 500 lb. per sq. in. of valve area +spring load.
Fuel Valve (a) Swedish type.
1000 lb. per sq. in. of needle area+spring
load+inertia, all reduced by the leverage
employed .
(b) Augsburg type.
Difference between the spring load and
1000 lb. per sq. in. of needle area at
stuffing-box.
Two Stroke Engines
Scavenge Valves. Spring load less scavenge air pressure
into area of valve +inertia of valve.
Fuel and Starting Valves. As for four stroke engines.
All the above are of course subject to sHght correction for
friction.
By way of example, the main dimensions of the exhaust
valve lever for a 20" four stroke cylinder
are calculated below : —
Data :
Diameter of exhaust valve, 6-5 in.
Fulcrum spindle and ex-
haust valve lever centres,
as in Fig. 241.
Pressure load on exhaust
valve
=0-785 X 6-52 X 45 = 1490 lb.
Spring load at, say, 8 lb.
per in.^ of valve area
=0-785x6-52x8= 2651b.
Inertia load, say . . 40 lb.
Total load to open ex-
haust valve
17951b.
Fig. 241.
Reaction at fulcrum spindle, about 3600 lb.
Bending moment at fulcrum spindle = —. in. lb.
Allowing a stress of 6000 Ib./in.^,
d^ 3600x6x18
24
10
6000 X 24
d=3in.
:2-7
272
DIESEL ENGINE DESIGN
Allowing for a bush J" thick and about |" metal at the boss
the external diameter of the latter will be 6".
Sketching in the approx-
imate outhne of the lever,
as in Fig. 242, it is seen
that at the weakest section
AA, the bending moment
is about 1800 X 18-5 in. lb.,
and taking a stress of
modulus Z of the section AA
Fig. 242.
5000
should be
lb. /in. for cast steel,
1800x18-5
5000
the
6-66 in. s
If the section AA is approximately T-shaped, as in Fig.
then Z=b.t.h. nearly, "h" is 5-75, and
therefore
243,
5-75
16 in.2
Fig. 243.
which is satisfied by b=2-25 and t =0-515".
The lever may be made of approximately
uniform strength by tapering towards the -g
ends, both in width and depth, as in Fig.
244, whilst the flange and web thicknesses
are kept constant. If a double bulb or
other section is required, for the sake of
appearance, and on casting considerations,
it is a simple matter to sketch in such a
section approximately equivalent to the
simple I-section to which the calculation
apphes.
Push-rods. — In some designs a push-rod is introduced
between the lever and the cam-roUer, as shown diagram-
matically in Fig. 199 ante, in order to enable the cam-shaft to
be located at a low level.
For this purpose bright
hoUow shafting, or even
black lap - welded steam
tubes are suitable, if not
too highly stressed.
For handy reference in designing such pnsh-rods the follow-
ing table, taken from Prof. Goodman's " Mechanics Applied to
Engineering," is given for the buckling loads of tubular struts.
In using these figures it is advisable to use a factor of safety of
Fig. 244.
VALVE GEAR
273
not less than about 3 or 4, and in no case to employ stresses
exceeding 10,000 Ib./in.*
Buckling Stress (Free Ends), Lb. per Sq. In.
Ratio l^"g**^-
Mild Steel.
diameter.
10
59,000
20
42,000
30
29,000
40
20,000
50
14,000
60
10,500
70
8,200
80
6,500
90
5,500
100
4,500
The jointed ends of the rods may be of forged steel bar or
malleable cast iron bushed with bronze, as in Fig. 245.
Cam-shafts. — ^In modern shops the cam-shaft may be
rapidly and cheaply ground to size from black bars. In order
to facilitate the driving on of the
cams for a multi-cylinder engine
the enlarged diameters are usually
made of increasing sizes, differing
by successive thirty-seconds of an
inch, or thereabouts, as shown
exaggerated in Fig. 246. The same
figure which represents the cam-
shaft for a four stroke generating
set of three cylinders also shows the
method of supporting the shaft by
means of a continuous trough with
one bearing between each bank of
cams. This arrangement has a very
neat appearance and makes pro-
vision for catching the oil which
drips off the cams and rollers. In
some designs the cams are allowed to dip into an oil-bath,
the level of which is maintained constant by a small pump
Fig. 245.
274
DIESEL ENGINE DESIGN
provided for the purpose, or by a connection taken from the
forced lubrication system. A copious supply of oil to the
cams has the advantage of securing quiet running.
In other designs the cam-shaft is supported by bearing
brackets secured to the cyHnders as in Fig. 247. In this case
it is very desirable to fit light cast or sheet iron guards round
each bank of cams. The cam-shaft bearings are divided
horizontally for adjustment, and the shells may be of cast iron
lined with white-metal, solid gun-metal, or in small engines,
where the cost of material does not outweigh the advantage of
simplicity, of solid die-cast white-metal. Owing to the slow
peripheral speed and the intermittent character of the loading.
G. 247.
grease lubrication by Stauffer boxes is quite adequate, although
ring and syphon are more commonly used.
A slightly different arrangement is shown in Fig. 248.
Here the shaft is supported by a series of cam-troughs, one to
each cylinder, each trough having two bearings. The extra
rigidity of this arrangement allows of the cam-shaft diameter
being reduced below the figure required with the other arrange-
VALVE GEAR
275
ments described. This division of the trough into segments is
advantageous from the manufacturing point of view as the
smaller parts are easier to cast and handle in the shops ; also
one pattern serves for engines of any desired number of
cylinders.
Strength of Cam-shafts. — The size of cam-shaft required for
a given engine would appear to depend not so much on the
stresses to which it will be subject, as on the rigidity necessary to
secure sweet running of the gear. For a four stroke engine the
opening of the exhaust valve against the terminal pressure in
the cylinder is the severest duty which the cam-shaft is called
Fie. 248
upon to perform. The load is applied and released fairly
suddenly, and a cam-shaft lacking in torsional and transverse
rigidity would undoubtedly be subject to oscillations, which in
an acute case would give rise to the following evils :—
(1) Noisy action of cams, due to torsional recoil of shaft
after each exhaust lift.
(2) Interference with the timing of valves (particularly the
fuel valves) of cylinders remote from the gearing end
of the cam-shaft.
(3) Chattering of the gear-wheels by which the shaft is
driven.
In view of the fact that as shaft diameters are increased the
stiffness increases at a greater rate than the strength, it seems
just possible that strength considerations might outweigh
those of stiffness in very large engines. On the other hand.
276
DIESEL ENGINE DESIGN
if angular deflection of the shaft between contiguous cylinders
be accepted as the criterion, then considerations of similitude
give shafts of diameters bearing a constant ratio to the cylinder
bores (or rather exhaust valve diameters) and constant stresses
in all sizes if the terminal pressure is always the same. The
fact that in practice relatively thinner cam-shafts are used in
large engines may be due to the lower terminal pressures
obtaining in the cylinders of the latter.
The following table shows the approximate diameters of
cam-shafts used in practice on four stroke engines of different
sizes : —
Bore of Cylinder in inches .
Diameter of Cam-shaft in inches
6 10 15 20
25 30
U 2i, 2| 3f
3i 4i
CamshaFt
The above figures hold for any number of cylinders up to
four with the cam-shaft drive at one end, or eight with the
cam-shaft drive at the centre.
For two stroke engines these diameters
may be materially reduced on account of
the absence of exhaust valves. Average
figures for existing practice appear to be
about 25% lower than those given above
for four stroke engines. The fuel pumps
and cylinder lubricating pumps are fre-
quently driven off the cam-shaft, but any
auxiliary gear, such as circulating pumps,
etc., requiring appreciable power are
precluded.
Cam-shaft Drives. — For non-reversible
engines, the spiral drive shown diagram-
matically in Fig. 249 is the favourite.
This drive comprises the following com-
ponents : —
(1) 4<ower spiral wheels.
(2) Footstep bearing for vertical shaft.
(3) Vertical shaft and couplings.
(4) Upper spiral wheels.
The lower spiral wheels generally have a
1 : 1 ratio, so that the vertical shaft runs
at engine speed. In some designs the
FiQ. 249. ratio is 1| : 1, and the vertical shaft runs
278
DIESEL ENGINE DESIGN
VALVE GEAR
279
cam-shaft, due to spiral wheels, etc., preferably by
ball-thrust washers.
(3) The wheels to run in a bath of oil, and suitable arrange-
ments to be made to prevent leakage of the latter.
(4) The general arrangement of gear-box and bearings to be
compact and in general conformity with the design of
the rest of the enginej
Probably the simplest way of fulfilling the above require-
ments is to cast the gear-case
en bloc, with a continuous cam-
trough, as in Fig. 246 ante, or
in the case of a central drive
to suspend the gear-case be-
tween two sectional troughs,
as in Fig. 254, by a sufficient
number of fitted bolts.
Spur and Spiral Gears for
Gam-shaft Drives. — The ques-
tion of the strength of the teeth
hardly arises in this case, and
the problem consists in the selection of materials and propor-
tions giving quiet running and absence of wear. The following
pairs of materials are in common use : —
Fig. 254.
Driver.
(1) Cast iron.
(2) Steel.
(3) Steel.
Follower.
Cast iron.
Cast iron.
Bronze.
Of these, the pairs (1) and (2) appear to give the best results,
with proper proportions and adequate lubrication, etc.
In good practice, the normal circular pitch of the teeth is
made about equal to one-twelfth of the cylinder bore for both
spur and spiral gears, and the width of face about one-fifth of
the cylinder bore in the case of four stroke engines. It is a
fairly safe rule to make the pitch as coarse as the smallest wheel
wiU allow in the case of spiral wheels. With spur wheels fine
pitches are not so objectionable as the sliding between the
teeth is much less.
For satisfactory running, the teeth must of course be
properly cut and the wheels accurately centred. The tooth
clearance should not exceed about 2/1000", and should be
uniform all round.
280
DIESEL ENGINE DESIGN
For particulars of tooth-gearing calculations the reader is
referred to the special books devoted to this subject. The
diagram shown in Fig. 255 is very useful in the preliminary
stages of spiral drive calculation. Suppose it is desired to
design a pair of right-angle spiral wheels of say 1 : 2 ratio ;
first calculate the diameters of a pair of spur gears of the
desired pitch and giving the desired ratio, viz., 1:2. Draw OA
and OB equal to the pitch radii of the follower and driver
respectively. Complete the rectangle OBCA and draw any
line DCE, cutting the axes in D and E. Then DC and CE will
be equal to the pitch radii of equivalent spiral wheels having
Fio. 256.
spiral angles a and /8 and having a normal pitch the same as
the circular pitch of the spur wheels first calculated. It usually
happens that the wheel centres (DE) are fixed within approxi-
mate limits by space considerations, and a process of trial and
error is required to find suitable values for the number of teeth
and the spiral angles. The l^ter should not be less than about
27°, as the efficiency falls o^rapidly as this figure is reduced.
Having obtained an approximate solution by the above
method, the angles should be determined to the nearest minute
by logarithmic trial and error calculation by means of the
following relation : —
AC , BC
sin /3 cos ^
=DE=required wheel centres.
VALVE GEAR
281
Reversing Gears. — In spite of early anticipations of diffi-
culty, reversing gears for Marine Diesel Engines have attained
a iiigh degree of efficiency. On the score of simplicity, relia-
bility and quick action they compare favourably with the
corresponding parts of steam engines. A very great number
of different gears have been suggested and patented, but those
in widespread use fall into two or three well-defined classes,
which will be described below.
Rollers Lifted Camshaft Slides, Rollers Replaced
Developed Surface oF Drum
Fig. 256.
Sliding Cam-shaft Type of Reversing Gear. — ^This type of
gear is the favourite for four stroke engines, though it is
equally applicable to those working on the two stroke cycle.
Ahead and astern cams side by side are provided for the opera-
tion of each valve. Reversal is effected by sliding the cam-
shaft a few inches endways in its bearings, so that the ahead
cam is removed from the action of the roller and replaced by
the astern cam and vice versa. It is in general necessary to
arrange means whereby the valve rollers may be swung clear
of the cam noses during the longitudinal movement of the
shaft, otherwise fouls would occur. In some very small
282
DIESEL ENGINE DESIGN
engines the necessity for such provision is obviated by employ-
ing curved-faced, rollers adapted to sMde up and down inclined
faces between the ahead and astern cams respectively.
The method adopted in some of the Burmeister & Wain
engines is shown in Fig. 256. A drum A is mounted on a
cranked shaft B, on which are hinged drag links C, connected
to the roller end of the valve push-rods D. Drum A is provided
with a groove E, the developed shape of which is shown in the
figure. This groove accommodates a roUer P, attached to a
movable collar bearing G. Shaft B is rotated in the direc-
tion desired (" ahead to astern " or " astern to ahead ") by
CamshaFt^s,
MuFr
Fig. 257.
suitable gearing in connection with a reversing servo-motor
or the hke. Approximately one-third of a revolution of the
shaft suffices to swing the rollers clear of the cams ; meanwhile
the cam-shaft is stationary. Another approximate one-third
of a revolution causes the groove E to shift the cam-shaft from
ahead to astern positions, or vice versa, whilst the valve rollers
execute a harmless movemCTit a little further out and back
again. The remainder of the revolution of the weigh-shaft B
replaces the rollers in their running position.
In other engines of the same make, the developed shape of
the groove E is executed on the back of a rack by means of
which the straight line motion of a vertical servo-motor is
converted into rotary motion of the weigh-shaft. This varia-
tion is shown in Fig. 257.
VALVE GEAR
283
If separate means be adopted for removing and replacing
the rollers, it is obviously possible to devise very simple means,
of shifting the cam-shaft as in Fig. 258 for example. In such
cases the two mechanisms must be interlocked to prevent a
false manoeuvre.
Experiments show that the force required to move the cam-
shaft longitudinally is about one-third of the weight of the
cam-shaft, plus cams and other gear keyed thereto, and this
figure may be used as a basis of calculation for this type of
gear. It is advisable, however, to allow a fair margin of power,
as the resistance to motion must always be a matter of some
uncertainty. When the axial motion of the shaft has the
^^^^^^^^^^^^
Fig. 259.
effect of opening one or more of the valves, the resistance dus
to this cause must be added to that of the shaft itself.
Twin Cam-shaft Type of Reversing Gear. — ^With this type
of gear, which is a speciality of the Werkspoor Company, not
only are separate cams provided for ahead and astern running,
but the latter are mounted on separate cam-shafts capable of
being slid into and out of action as required. Fig. 259 illus-
trates the arrangement diagrammatically. The cam-shaft
drive is usually by means of coupling rods. The chief advan-
tage of this gear would appear to be the absence of special gear
for swinging the rollers out of operation, this process being un-
necessary. It is perhaps worth noticing here that the sim-
plicity of the manoeuvring gear is one of the outstanding
features of the Werkspoor Marine Diesel Engine.
Twin Roller Type of Reversing Gear. — This gear depends
on some form of link-work such as that shown in Fig. 260.
Rollers A and B lie in the planes of the ahead and astern cams
respectively. In the position shown the timing of the valve
284
DIESEL ENGINE DESIGN
is controlled by the ahead cam and roller A ; roller B is mean-
while outside the radius of action of its cam. Rotation of the
weigh-shaft C through a predetermined angle throws roUer A
out of action and brings roUer B into action with the astern
cam. This type of gear has been applied to both four and two
stroke engines.
Ahea'd Cam' ^Astern Cam
Fig. 260.
A different arrangement, having some slight resemblance to
the above, is shown in Fig. 261. In this case there is only one
roller which is swung from the ahead to the astern cam by a
motion in a plane at right angles to the plane of the gear. The
roUer face is curved to allow of this slight angular displace-
ment from the vertical. The inherent defects of this mechanism
probably render it unsuitable for use in conjunction with any
but the air starting valves.
Selective Wedge Type of Reversing Gear. — This ingenious
gear, illustrated diagrammatically in Fig. 262, has been devised
by Carels Fr^res and used in connection with the starting air
and fuel valves of two strokamarine engines designed by them.
Ahead and astern cams are ^ovided side by side, and the valve
roUer A is wide enough to cover both. Between the cams and
the lever is interposed a roller-wedge piece B, under control of
a cam 0, mounted on a manoeuvring shaft D. The latter is
capable of independent rotary and endway motion. A suit-
able rotary motion of the shaft D withdraws the wedge B to
an extent which renders inoperative the ahead cam on which
it rests. A longitudinal movement of shaft D carries the
VALVE GEAR
285
286
DIESEL ENGINE DESIGN
wedge B with it, and a further rotation of D introduces the
wedge between roller A and the astern cam, and vice versa
for astern to ahead. It is to be noticed that by a suitable
arrangement of the durations and sequences of the cams by
which the wedges are operated the engine is caused to start up
in any predetermined manner, as for example : —
Position (1) Six cylinders on air. Fuel valves inoperative.
,, (2) Three cylinders on air. Three cyKnders on fuel.
,, (3) Six cylinders on fuel. Air valves inoperative.
Special Reversing Gears for Two Stroke Engines. — ^With
two stroke engines there are a number of means by which the
duplication of cams may be avoided. Considering the case of
Astern-
Fuel Injection.
PeriooAstern
Ahead
Fuel Injection
Period Ahead
Scavenae Period,
Mead
Open Ahead
CloseAstem
Scavenge Period
Astern
an engine fitted with scavenge valves and neglecting the start-
ing air valves for the moment, the valve settings for ahead and
astern will be somewhat as shown in Fig. 263. It wiU be seen
that for both fuel and scavenge valves all that is required to
effect reversal is the rotatron of the cam-shaft through a
certain angle a for the fuel valve and /8 for the scavenge valves.
In some early engines it was decided to select a=(8=about 30°
to 35°, and so effect reversal of both valves by one movement.
In later engines, however, it is more usual to use the rotation
of the cam-shaft to reverse the scavenge valve only and adopt
independent means, such as duplicate cams, etc., for the fuel
and starting valves. The effect of the rotation of the cam-
VALVE GEAR
287
shaft on the settings of the latter must of course be allowed for
in fixing the angular positions of the fuel and starting cams.
A simple method of effecting the desired rotation of the
cam-shaft of a small engine is shown diagrammatically in
Fig. 264. Spiral drives are used and the vertical shaft is in
two pieces, C and D, connected by a splined coupling permit-
ting vertical movement of the upper half D. The vertical
Camshaft
Camshaft
Jll / ^mL^CrankshaFt
Fig. 264.
'rankshaft
Fio. 265.
position of D is determined by the lever E, which is hinged at P
and connected at G to an eccentric or other suitable means of
transmitting motion from the hand- wheel. The extreme upper
and lower positions of D determine the ahead and astern
running positions.
If h=Lift of vertical shaft,
r= Pitch radius of wheel B,
Then - = Reversing angle in radians.
288 DIESEL ENGINE DESIGN
In other arrangements the vertical shaft is moved as a whole,
as in Fig. 265, and in calculating the amount of motion required
for a given reversing angle it is necessary to take into account
the rotation of the vertical shaft due to the sliding between the
lower helical wheels.
Consider the case where both upper and lower gears have a
ratio of 1 : 1 . Let a be the spiral angle of the crank-shaft gear-
wheel.
Note that a<45°.
Let h =Lift of vertical shaft.
ri=Pitch radius of vertical shaft lower wheel.
r2=Pitch radius of cam-shaft wheel.
Then
Rotation of cam-shaft due to axial movement of vertical
shaft =— radians as before.
1-2
Further,
Rotation of vertical shaft due to sliding of lower spiral
. , h.tan a
wheels =
Therefore,
T, . , h htana
Keversmg angle = -±
With the arrangement shown the positive sign applies when
the upper and lower spirals have the same hand and the
negative sign when they are of opposite hand.
An inspection of the valve settings for ahead and astern,
as shown in Fig. 263 ante, reveals the fact that the " reversing
angle " is always described in the direction opposite to the
previous direction of motion. Advantage has been taken of
this fact to obtain self-reversing valve settings by arranging,
between the cam-shaft drive and the cam shaft proper, a
claw clutch having angular clearance between the jaws equal
to the reversing angle, ^ith this arrangement independent
reversible gearing must be used for the starting air valves.
A suggested improvement on the above is to provide mechanical
means for taking up the slack between the jaws whilst the
engine is standing, instead of allowing it to be suddenly taken
up on starting.
Movable Roller Type of Reversing Gears. — Instead of
rotating the cam-shaft through a certain angle relative to the
VALVE GEAR
289
cam-roller, the same effect may be obtained by turning the
roller relative to the cam-shaft.
One such arrangement is shown in Fig. 266. It will be
noticed that the valve lift is less in the astern position than in
the ahead, but this is unimportant. A similar device is shown
in Fig. 267. In both these designs the reversing angle is
conveniently halved by fitting double-nosed cams to a half-
speed cam-shaft.
Another gear coming under this category is shown in Fig. 268.
'Valve Lever
■Ahead
Astern
The displacement of the roller from its ahead to astern position
is effected by the partial rotation of the eccentric fulcrum A,
and the roller passes through a neutral position, in which it is
outside the radius of operation of the cam.
The above descriptions by no means exhaust the list of
existing Diesel Engine reverse gears, and doubtless others
remain to be invented. It is evident, therefore, that the
problem of reversibility no longer presents any obstacle to the
development of the Diesel Engine for marine service.
For slow-running engines there is probably little serious
objection to any of the gears which have been described For
high speed engines, however, most existing gears for two
290
DIESEL ENGINE DESIGN
stroke engines are noisy and of doubtful durability. Curiously
enough, the sliding cam-shaft scheme, which would seem to
possess every advantage, does not appear to have been used
on this class of engine so far.*
Astern
Fig. 267.
Manoeuvring Gears. — In this connection the term manoeuv-
ring gear is applied to those mechanisms apart from the
reversing gear which come into operation on starting up a
marine engine.
* This gear has recently been adopted by the firm of Franco Tosi for their
two stroke engines.
VALVE GEAR
291
The procedure differs in different designs, but in general the
following remarks are applicable : —
(1) When the engine is standing the blast air supply should
be cut off, to prevent accumulation of pressure in any cylinder
the fuel valve of which happens to be open. If means be
provided for putting the fuel valves out of operation in the
stop position the blast cut-out is not so essential, but is still
desirable as a safeguard.
(2) The blast air should be turned on automatically im-
mediately the engine is started, although it is quite advan-
tageous to provide an independent shut-off and regulating
valve under the control of the engineer.
Fio. 268.
(3) When the engine is standing the starting air should be
cut o£E, as there is otherwise great loss of air due to leakage
past the starting valves. The starting air shut-off may be
automatic or hand operated ; if the latter, it should be opened
and closed by one simple motion. An ordinary high pressure
globe valve, fitted with a quick-threaded spindle, is suitable
for this duty.
(4) The fuel pump suction valves should be held off their
seats until such time as the fuel valves are in running position,
independent of the position of the fuel control.
(5) The fuel control should be a handle (not a wheel) with
a wide range of movement between no oil and full oil.
(6) A wheel, or better stiU a lever, is provided in connection
with suitable mechanism for putting the starting valves into
operation at starting, and subsequently putting them out of
292 DIESEL ENGINE DESIGN
operation when sufficient speed has been attained to ensure
firing in the cylinders. The same mechanism may or may not
(in different designs) throw the fuel valve mechanism out of
and into operation. Furthermore, in some designs the opera-
tion of this gear is graduated, as in the foUoAving scheme,
which refers to a six-cylinder engine :^
First notch. — Six cylinders on air (starting).
Second notch. — Three cylinders on air. Three cylinders on
fuel.
Third notch. — Six cylinders on fuel.
It has been the experience of many engineers that this
graduated control, besides adding to the complication of the
gear, is a positive drawback, and that the greatest certainty of
starting is secured by passing direct from all cylinders on air
to all cylinders on fuel. This arrangement has been adopted
on many large four stroke engines with every success, and
appears likely to become standard practice. The gear under
consideration is connected with the fuel pumps and with the
blast starting air supply, so that the following conditions are
secured : —
(a) Movement of the lever towards the starting position
automatically turns on the starting and blast air.
(b) Suction valves of all fuel pumps held off their seats.
Further movement of the lever puts the starting valves of
some or all of the cylinders out of operation, and simultan-
eously allows normal operation of the corresponding fuel
pumps and also of the fuel valves if these latter are arranged
to be out of operation during the time the starting valves are
working.
(7) Some simple type of interlocking gear is usually fitted
to prevent the following false manoeuvres : —
(a) Starting the engine before the reversing gear is in either
the full ahead or fulliistern position.
(b) Operating the reversing gear before the manoeuvring
gear has been put into the stop position.
Some of the means adopted to secure the conditions out-
lined in sections (1) to (7) will now be described. Further
reference need not be made to the reversing gear, as with the
exception of the interlocking arrangements mentioned above
VALVE GEAR
293
the reversing arrangements are entirely independent of the
manoeuvring gear.
A simple type of manoeuvring gear is shown diagram-
matically in Fig. 269. The fuel and starting levers are eccen-
trically mounted on fulcrum shafts, as
described earlier in this chapter. Each
fulcrum shaft A is connected by links
and levers to a manoeuvring shaft B,
under the control of a hand lever C.
In the upper position of lever C aU the
fuel valves are in operation, and in
the lower position the starting air
valves. A link D connects the
manoeuvring lever to an eccentric
fulcrum on the fuel pump, by means
of which the suction valves are lifted
by suitable tappets provided for this
purpose, during such time as the
starting air valves are in operation.
Another link performs a similar opera-
tion on the blast air control valve, but
in this case the connection is such that
the blast air is only cut off in the
neutral or stop position of the
manoeuvring lever.
In some engines the above arrangements are adopted in
principle, but two separate control gears and levers are pro-
vided for the forward and aft halves of the engine. The two
control levers are placed close together, so that the engineer
can work one with either hand. On starting he pulls both
towards him, thus putting all cylinders under starting air.
As soon as sufficient speed has in his judgment been attained,
he pushes one lever towards the fuel position. If firing starts,
he then pushes the other lever into the fuel position. If on the
other hand firing does not ensue, he may pull back the lever
into the starting air position and try the other lever in the fuel
notch. With an engine in good order it is probably advan-
tageous to throw over both levers simultaneously.
In other designs it is not necessary to operate the fulcrum
shafts, as the starting valves automatically throw themselves
into operation when starting air is turned on, and become in-
operative when the starting air pressure is released.
Fig. 269.
294 DIESEL ENGINE DESIGN
A great deal of ingenuity has been expended on the design
of gears for throwing successive combinations of cylinders
from " air " to " fuel " positions by a continuous movement
of a wheel. Some designers have even gone the length of
combining the reversing and manoeuvring mechanisms so that
aU positions ahead and astern are secured by clock-wise and
anti-clock- wise rotation of this wheel. In the writer's opinion,
such gears are not to be desired, for the following reasons : —
(1) Intelligent manipulation of machinery involves a certain
parallelism between the mental state of the operator and the
response which the machine makes to his control. This state
would appear to be most easily secured when separate and
distinct operations on the part of the machine are made in
response to separate and distinct movements on the part of the
operator. Hence it would appear best to keep the reversing
and manoeuvring control separate, with the exception of what-
ever measure of interlocking is necessary to prevent accidents.
(2) Gears of the kind referred to are usually complicated,
and not easily understood or overhauled.
(3) The complication of such gears is not infrequently
associated with backlash, which renders accurate valve setting
difficult to effect and maintain.
(4) Complicated gears do not appear to have any practical
advantages to offset their increased cost.
Interlocking Gears. — The precise form which an interlocking
gear takes in any design depends on the forms of mechanism
adopted for the reversing and manceuvring gears respectively,
but the problem very frequently reduces to that of two shafts,
either of which shall only be capable of movement in prescribed
positions of the latter. A simple interlock for two parallel
shafts, subject to partial rotation, is shown in Fig. 270. It will
be observed that the manceuvring shaft A can only be rotated
when the reversing shaft B i^in one of two positions (ahead
and astern) defined by the positions of the gaps cut in the
circumference of a disc keyed thereto. Furthermore, the
shaft B can only be rotated from its ahead to its astern position
(or vice versa) when the manceuvring shaft A is in one position
— ^the stop position. The solution when the shafts are at right
angles, as in Fig. 271, is equally obvious. An indefinite
number of other schemes could easily be devised to meet the
requirements of different arrangements of gear.
VALVE GEAR
295
Hand Controls. — It is essential that all the wheels and
levers by means of which the engine is controlled should be
grouped together so that they may be manipulated by one
man in one position. In the Werkspoor Engine the arrange-
ment of controls is particularly neat, and consists of a row of
hand levers, arranged at a convenient height at the centre of
the engine, and one of which is used to answer the telegraph.
In the Burmeister & Wain Engines one long hand lever is
Fig. 270.
Fig. 271.
provided for the manoeuvring process of switching over from
air to oil, and also the regulation of the fuel supply. A separate,
smaller hand lever is provided for reversing, and a wheel is
fitted for emergency hand reversing. Under normal conditions
the engine is therefore controlled by two levers.
In a recent Polar Diesel Marine Engine a somewhat novel
departure has been made in centrahsing the controlling gear
on the top platform. This position gives the engineer an
advantage in having the fuel valves under observation. The
fuel pumps are also grouped in the immediate neighbourhood.
INDEX
Adiabatic expansion and compres-
sion, 18
After burning, 6, 28
Air-bottles, 252
— composition of, 34
— compressors, 244
cranks, 73
drives, 245
valves, 247
— -motors, 261
— specific heats of, 18
— • starting valves, 257
— suction pipes, 220
valves, 222
Alternators in parallel, 104
Balance sheet, heat, 27
Bearings, cam-shaft, 274
— lubrication of, 127
— main, 126
— Mitchell thrust, 26
— shells for, 128
Bedplates, 125, 132
Bending in connecting rods, 181
et eeq.
— in crank-shafts, 77 ei eeq.
— in valve levers, 271
Blast air, functions of, 243
pipe system, 254
stop valve, 255
Bore and stroke, determination of, 59
Bottles, air, 252, 253
Burmeister & Wain marine engines,
14
fuel valve, 216
reversing gear, 282
starting valve, 258
Cams, 263, 264
— rollers for, 266
Cam-shafts, 273
— diameter of, 276
_ drives for, 131, 132, 276, 279
Cam-shafts, strength of, 275
Chalkley, 15, 36
Charge, renewal of air, 37
Clerk, 36
Columns " A " type, 119, 120, 121
Combustion stroke, 5
Compression, adiabatic, 18
— isothermal, 17
— polytropic, 19
— pressure, 2
effect of high, 3
— ■ stroke, 5
— temperature, 3
Compressors, air, 244 et seq.
— ■ calculation for, 248
— scavenge air, 237
— three and two stage, 250
Connecting rods, 176
bending in, 181
big ends of, 177, 178
— — bolts for, 188
example of calculation for, 185
material for, 176
proportions of, 187
Couplings, crank-shaft, 72, 96
— vertical shaft, 277
Cranks, air compressor, 73
— scavenger, 74
Crank-cases, 136
— thickness of, 139
— type of framework, 121
Crank-shafts, bending in, 77 et seq.
— ■ construction of, 65
— couplings for, 72, 96
— lubrication of, 69, 70
— material for, 65
— proportions of, 75
— stresses in, 95
— twisting in, 89
— webs of, 70
Critical speeds, 108 et seq.
Crossheads and guides, 121, 174
influence on lubrication, 69
pressure diagrams, 175
Cylinder covers, 152 et seq.
defects of, 154
for two stroke engines, 158
297
298
DIESEL ENGINE DESIGN
Cylinder covers, proportions of, 156
strength of, 156
— jackets, 145 et aeq.
— liners, 143, 144
— lubrication, 151
— types of, 141 et aeq.
D
Dalby, 36, 176, 251
Diagrams, crank angle, 43
— entropy, 28
— • indicator, for four stroke engine,
6, 7, 20, 32, 91
for two stroke engine, 18
— of port and valve areas, 44
— twisting moment, 93, 101, 102, 103
— valve areas, 42
Diesel Engine, ideal, 19
types of, 12
— • principle, 1
Dowels, 116
E
Efficiency of combustion, 28
— mechanical, 25, 27
— thermal, 16 ef seq.
— ■ volumetric, of air compressors, 248
of cylinders, 7
Energy absorbed by fly-wheels, 99
Entropy diagrams, 28 et seq.
Exhaust gas, specific heat of, 34
— • process, calculation for, 47 et seq.
in two stroke engines, 46
— stroke of four stroke engine, 6
— system, 241
Expansion, adiabatio, 18
— isothermal, 17
— polytropio, 19
F
Flame plate, 212
Fly-wheels, 97 et seq.
— effect on critical speeds of, 108
defined, 97 %
on momentary governing of,
100
— energy absorbed by, 99
— example of design of, 118
— for alternators in parallel, 104
— ftmotions of, 97
— radius of gyration of, 113
— strength of, 115
— ■ types of, 113
Ford, 262
Foui stroke cycle, 4
engines, 12, 13, 14
Framework, 119 ei seq.
— " A " frame type, 119
— • crank- case type, 121
— machining of, 140
— ■ staybolt type, 125
— trestle type, 123
Frequency of oscillations, 109
Fuel, calorific value of, 16
— consumptions of actual engine,
24
^ at var3ring load, 24
^ of ideal engine, 24
tables of, 27
— distributors, 193
— -injection valves, 211 et seq.
Augsburg type, 212
Burmeister, 216
— — — closed, 212
— Swedish, 214
— -piping, 193
— - pumps, 194 et seq.
calculation for, 200
details of, 196, 200
— . — .plungers, 203
valves, 203
— system on engine, 192
external, 190
Funok, 52
G
Gases, exhaust, discharge of, 6, 46
et seq.
specific heat of, 34
— flow of, through orifices, 37
Gear, barring or turning, 97
— indicating, 71, 188
— manoeuvring, 290
— - reversing, 281 et seq.
— ^ running, 161 ei aeq.
— spiral, 131, 276, 280
— spur, 279
— - valve, 263 et seq.
Girders, main bearing, 128
— side, of bedplate, 120
Goodman, 242
Governing, modem refinements of,
210
— -momentary, 100
Governors, 206 et seq.
— design of, 208
— sleeve type, 209
— springs for, 207
Gudgeon pins, 166
bearings for, 178
INDEX
299
Gudgeon pins, location of, 165
• lubrication of, 166
Guest, 76
H
Hand controls, 295
Heat balance sheet, 27
— ■ specific, of air, 18
-of exhaust gases, 34
— ■ — -of various gases, 34
— transmission, 250
Hurst, 160
Ideal engine, 19
Ignition temperature, 2
Indicated efficiency, 23
— ■ mean pressure, 23, 24, 32, 54
variation with bore, 63
— volumetric efficiency, 7
Indicating gear, 188
Indicator diagrams, four cycle, 6, 7,
20, 32, 91
— two cycle, 18
Inertia of reciprocating parts, 79,
172, 180
— of valves and gearing, 230 et seq.
Injection of air (See blast air)
— of fuel, 2
— • solid, 1
— valves (see fuel valves)
Inlet valves, 221
Interference of exhaust, 7
Interlocking gears, 294
Isothermal expansion and compres-
sion, 17
Joules' equivalent, 16
Judge, 36, 251
Levers, valve, 266
Liners, cylinder, 142-144
Literature, references to, 15, 36, 52,
64, 118, 140, 160, 189, 219, 242,
251, 262
Lubrication of cam-shaft bearings,
274
— of orank-pins, 69, 70
— of cylinders, 151
— • of fuel pumps, 204
— of gugdeon pins, 166
— of main bearings, 127
— . of (roller pins, 266 j
_ of jvalve levers,»270
M
Manceuvring gears, 290
Marine engines, 14
Mean indicated pressure, 23, 24, 32,
^*
Mechanical efficiency, influence of
size on, 26
— — table of, 27
— losses, 25
Mellanby, 118
Mexican fuel oil, 204
Milton, 15
Mitchell bearings, 26
Momentary governing, 100
Morley, 118
N
Nicholson, 251
Nitrogen, specific heat of, 34
Node, 109
O
Oil, drainage of lubrication, 133
Oil fuel, Mexican crude, 204
— — storage and filters, 190
■ system, external, 190
tar, 2, 216, 219
Order of firing, 66
Oscillations, torsional, 108 et seq.
Oxygen, specific heat of, 34
Petavel, 251
Petter, 52
Pilot ignition, 210
Piping, blast air, 254
— fuel, 192, 194
— starting air, 256
Pistons, 161 et seq.
— cooUng, 169, 170
— cores for, 164
— cracks in, 162, 163
— crowns for, 163
— for crosshead engines, 169-172
— pins (see gudgeon pins)
— rings, 164
— rods, 172, 173
— speeds, 60-62
— trunk, 161-169
Plvmgers, air compressors, 246
— fuel pump, 200
Polytropic expansion and compres-
sion, 19
Porter, 219
300
DIESEL ENGINE DESIGN
Pressure, blast air, 244
— compression, 2
— main bearing, 127
— maximum cycle, 2
— scavenge air, 44, 237
— teiminal, 6
Pulverisers, 213, 214
Pumps, fuel, 194 et seq.
— lubricating, 151
— scavenge, 237
Push-rods, 272
— buckling loads of, 273
B
Reliability, effect of incomplete com-
bustion on, 28, 243
of high cycle temperature on, 243
Renolach, 219
Reservoirs, air, 251, 254
Reversing gears, 281 et seq.
for two stroke engines, 286
— ■ — moving roller type of, 288
selective wedge type of, 284
sliding cam-shaft type of, 281
twin cam-shaft type of, 283
roller type of, 283
Revolving parts, centrifugal force of,
79
fly-wheel effect of, 90, 112
Richardson, 140, 160
Rubbing speed of main bearings, 127
Scavenge air pressure, 44, 237
receivers, 238
temperature, 41
valves, 239, 240
velocity, 41
— controlled port, 11
— process described, 9, 41
— simple port, 10
Scavengers or scavenge pumps, 237
Semi-Deisel engines, 1
Shafting, torsional oscillations of, ^^8
et seq. "
Silencers, 236, 242
Similar engines defined, 53
properties of, 54, 55
relative weight of, 54
Similitude, principle of, 53 et seq.
Smith, 76, 115, 160, 164, 187, 219
Solid injection, 1, 243
Specific heats of air, 18
of o^her gases, 34
Springs formulse, 234
Stationary engines, 12, 13
M.I.P. for, 62
piston speeds of, 62
weight of, 56
Studs, table of, 130
Suction pipes, 220
— pressure and velocity, 39, 40
— stroke of four cycle engine, 4, 39
— valves for engine, 222
for fuel pump, 203
Supino, Bremuer & Richardson, 15
Surface ignition engines, 1
Tar oil, 2, 216, 219
Temperature change during poly-
tropic process, 19
— for ignition, 2
— stresses in cylinder covers, 154
in piston crowns, 162
Thermal efficiency, \& et seq.
Torsional oscillations, 108 et seq.
Trestle type of framework, 123
Two stroke engines, 8, 9
exhaust and scavenge of, 8,
41 et seq.
— ■ fuel consumption of, 27
horizontal, 1 3
marine, 14
mean indicated pressure of,
62
mechanical efficiency of, 27
U
Unwin, 118
Value, calorific, 16
Valve-controlled port scavenge, 11
— gear, "238 6* seq.
— setting diagrams for four stroke
engine, 8
for two stroke engine, 9
Valves, air bottle, 253
suction, 221
— blast air shut off, 255
— compressor, 246
— exhaust, 223 et seq.
casings for, 226, 227
guide ends for, 225
heads for, 223
inertia of, 230
lifting devices for, 226
springs for, 229, 234
INDEX
301
Valves, fuel injection, 211 e< seq.
Augsburg type, 212, 218
Burmeister type, 216
closed type, 212
open type, 211
Swedish type, 214, 218
pump, 200
— inlet (see " Air suction valve ")
— levers, 266 et seq.
— needle, 212, 213
— starting air, 256 et seq.
■ Burmeister & Wain type, 258
Velocity of exhaust gases, 48, 235
— of scavenge air, 41
Velocity of suction air, 40, 223
Volumetric efficiency of air com-
pressors, 248
of engine cylinders, 7
W
Water-cooled pistons, 169, 170
Webs of cranks, 70
Weights of engines, 56 et seq.
Wells and Taylor, 15
Wimperis, 36
Work done by expansion of a gas,
17, 19
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